GIFT  OF 


Civil  Eng. 


UNIVERSITY 


.  .  •        -,  -  ,  ..'•<• 

.r» 


jri'-'  '•" 


BERKEUEY. 

ENGINEERING    LIBRARY 
OF 

WILLIAM   B.   STOREY 

A   GRADUATE  OF 

THE    COLLEGE    OF    MECHANICS 
CLASS   OF  1881 

PRESENTED  TO  THE  UNIVERSITY 
1922 


MODERN 


LOCOMOTIVE  CONSTRUCTION 


BY 


J.  G.  A.  MEYER 


ASSOCIATE  EDITOR  OF  THE  "  AMERICAN  MACHINIST  ; "   MEMBER  AMERICAN  SOCIETY  MECHANICAL 
ENGINEERS  ;  FORMERLY  CHIEF  DRAFTSMAN  AT  THE  GRANT  LOCOMOTIVE  WORKS 


Illustrated 


NEW   YORK 
JOHN    WILEY   AND   SONS 

53   EAST   TENTH   STREET 
1892 


IV 


I'll E  FACE. 


The  principles  upon  which  the  methods  of  construction  are  based  are  generally 
given,  so  as  to  enable  any  one  who  designs  with  the  necessary  regard  to  theory  to 
make  modifications  to  suit  his  own  ideas.  In  writing  the  original  articles  I  have  aimed 
to  make  each  one  as  complete  as  possible.  This  necessarily  required  a  few  repetitions, 
which  were  allowed  to  remain  in  this  book,  as  it  was  thought  such  a  course  would 
make  the  book  more  convenient  for  reference. 

My  thanks  are  due  to  Mr.  P.  Arnot,  Supt.  of  the  Grant  Locomotive  Works,  when 
these  works  were  located  at  Paterson,  N.  J. ;  to  Mr.  John  Headden,  formerly  Supt.  of 
the  Roger  Locomotive  Works ;  Mr.  D.  Shirrell,  draftsman  at  the  Richmond  Locomo- 
tive Works ;  and  other  friends,  for  valuable  suggestions  and  assistance  in  the  prepara- 
tion of  this  work.  I  am  also  indebted  to  Mr.  Theo.  N.  Ely,  General  Supt.  of  Motive 
Power,  Pennsylvania  R.  R. ;  the  Grant  Locomotive  Works ;  the  Cooke  Locomotive 
Works,  of  Paterson,  N.  J. ;  the  Rogers  Locomotive  Works,  of  Paterson,  N.  J. ;  the 
Rhode  Island  Locomotive  Works,  Providence,  R.  I. ;  the  Baldwin  Works,  Philadelphia, 
Pa. ;  the  Richmond  Locomotive  and  Machine  Works,  Richmond,  Va. ;  and  Mr.  A.  J. 
Pitkins,  Supt.  of  the  Schenectady  Locomotive  Works,  Schenectady,  N.  Y.,  for  kindly 
furnishing  me  with  drawings  and  data. 

PATERSON,  N.  J.,  September,  1892. 


CONTENTS. 


CHAPTER  I. 

PAOE 

INTRODUCTORY  REMARKS. — CLASSIFICATION  OP  LOCOMOTIVES. — TRAIN  RESISTANCE. — TRACTIVE 
POWER. — WEIGHT  OF  ENGINES 1 

CHAPTER  II. 

CONSTRUCTION  OF  CYLINDERS. — STEAM  PIPES. — SLIDE  VALVES 20 

CHAPTER  III. 
VALVE  GEAR. — CONSTRUCTION  OF  LINKS 74 

CHAPTER  IV. 

PISTONS. — CROSSHEADS. — SLIDES. — STUFFING  BOXES 138 

CHAPTER  V. 
FRAMES  AND  PEDESTALS. — AXLE  BOXES 182 

CHAPTER  VI. 
DRIVING  AXLES. — DRIVING  WHEELS. — COUNTERBALANCE 214 

CHAPTER  VII. 
MAIN-RODS. — SIDE-RODS. — CRANK-PINS 267 

CHAPTER  VIII. 

THROTTLE  PIPES. — THROTTLE  VALVK  GEAR. — SAFETY  VALVES. — WHISTLE.— PI-.MI-S. — CHECK- 
VALVES  343 

CHAPTER  IX. 
SPRING  GEAR  AND  SPRINGS 400 

CHAPTER  X. 

BOILERS. — GRATE  SURFACE.— HEATI MI  SURFACE. — RIVETED  JOINTS. — EXTENSION  FUONTS...  417 


vi  CONTEXTS. 

CHAPTER  XI.  |>A,.E 

ASH-PANS. — SMOKE-STACKS. — EXHAUST-PIPES -101 

CHAPTER  XII. 

SAND-BOXES. — BELLS. — PILOTS. — ENGINE-BRACES 510 

CHAPTER  XIII. 
ENGINE  TRUCKS 535 

CHAPTER  XIV. 
OIL-CUPS. — VALVES. — COCKS. — INJECTOR 552 

CHAPTER  XV. 
TENDERS. — TENDER-TRUCKS 563 

CHAPTER  XVI. 
USEFUL  RULES,  FORMULAS,  AND  DATA 595 

CHAPTER  XVII. 
COMPOUND  LOCOMOTIVES 617 


MODERN 

LOCOMOTIVE    CONSTRUCTION. 


CHAPTER    I. 

INTRODUCTORY  REMARKS.— CLASSIFICATION  OF  LOCOMOTIVES.— TRAIN  RESIST- 
ANCE.—TRACTIVE  POWER.— WEIGHT  OF  ENGINES. 

1.  Ill  late  yours  a  change  in  the  management  and  treatment  of  the  locomotive  lias 
taken  place  on  most  of  our  principal  railroads.  This  change  necessarily  caused,  also,  a 
change  in  the  construction  of  the  locomotive,  besides  other  improvements  that  have 
been  added  from  time  to  time. 

"2.  It  is  the  writer's  intention  to  give  in  these  chapters  a  general  description  of  the 
principal  parts  of  the  modern  locomotive,  illustrated  by  good  and  correct  drawings, 
and  indicate  the  improvements  that  have  been  added. 

Since  all  the  illustrations  represent  separate  parts  of  the  modern  locomotive,  and 
since  some  of  them  will  be  arranged  in  a  tabular  form  (an  arrangement  the  writer  has 
not  seen  in  any  book,  and  consequently  believes  it  to  be  new),  we  trust  that  these 
illustrations  will  be  appreciated  by  the  professional  designer,  and  by  the  young  designer 
in  particular. 

3.  In  order  to  make  the  reading  of  these  papers  profitable  to  the  mechanic,  we 
will,  in  connection  with  tin;  illustrations,  give  rules  relating  to  the  proportioning  of  tin- 
parts  in  as  plain  and  simple  language  as  we  can  command,  so  that  any  one  engaged  in 
the  building  and  running  of  tin-  locomotive  may  easily  understand  these  rules.  \\Y 
also  hope  that  the  description  of  the  locomotive,  which  is  almost  inseparable  from  a 
subject  presented  in  a  manner  as  we  propose  to  do,  will  prove  interesting  to  the 
ordinary  reader. 

Should  any  of  our  professional  friends  pronounce  the*  rules  given  as  something 
superfluous,  because  they  may  be  found  in  the  many  excellent  books  already  published, 
or  should  .-my  of  our  friends  find  fault  with  the  practical  method  of  treating  this 
subject,  we  would  kindly  remind  them  that  these  chapters  are  intended  fora  large  class 
of  readers — for  the  mechanic  and  engineer  in  particular — and  not  fora  favored  few. 


,'•  *'."•' t    i  .'.  '  '•. 

vU  :  "'•',  :  -•• 

I,'  <  *  l'     «»»««» 


MODERX  LOCOMO  Tl  1 7v    f  'O.Y.S  Tit  rt '  TION. 


4.  When  a  comparison  is  made  between  the  locomotives  built  recently  and  the 
locomotives  in  use  about  ten  or  twelve  years  ago,  a  change  in  their  construction  and 
appearance  will  be  noticed.     This  change  is  due  to  the  desire  of  railroad  managers  to 
reduce  the  cost  of  transportation  of  passengers  and  freight,  and  to  a  great  extent  this 
desire  has  been  realized. 

5.  In  former  times  it  has  been  the  custom  to  place  an  engine  in  the  hands  of 
one  engineer,  and  whenever  the  engine  was  attached  to  a  train  this  same  engineer 
had  his  hand  on  the  throttle  lever.     When  the  trip  was  completed,  the  engine  was 
carefully  housed  and  cleaned,  and,  so  to  speak,  was  put  to  rest.     In  fact,  we  once  heard 
an  engineer  say  (and  we  have  reason  to  believe  that  he  was  in  earnest)  that  engines 
needed  rest  as  well  as  engineers,  because  he  noticed  that  his  engine  never  worked  as 
well  when  Hearing  the  end  of  a  trip  as  it  did  when  starting.     If  this  engineer  is 
still  among  the  living,  he  must  either  have  changed  his  opinion  or  stepped  off  the 
footboard  to  stay  off.     The  trips  also  were  comparatively  short,  and  generally  the 
trains  comparatively  light.     Indeed,  when  we  carefully  consider  the  management  of 
the  locomotive  and  the  treatment  it  received  in  former  times,  we  might  almost  conclude 
that  the  locomotive  was  looked  upon  as  a  delicate  piece  of  machinery  that  needed 
extraordinary  care  to  keep  it  in  good  working  order. 

But  now,  mark  the  change  of  treatment  of  the  engine  that  has  taken  place  on 
some  of  our  best  railroads.  Notice,  for  instance,  on  these  roads  the  modern  freight 
locomotive  as  it  starts  off  with  as  heavy  a  train  as  it  can  possibly  haul  on  a  trip  of 
great  length,  the  engineers  relieving  each  other  at  designated  stations,  instead  of  one 
engineer  having  charge  of  the  engine  during  the  whole  trip,  as  in  former  times;  notice 
also  the  scanty  accommodations,  if  any,  for  cleaning  or  housing  the  engine  when  the 
trip  is  completed ;  the  short  time  the  engine  is  allowed  to  stand  still  after  it  has  been 
examined  and  found  to  be  in  good  working  order ;  the  heavy  train  it  must  haul  on  the 
homeward  journey,  run  by  any  engineer  that  is  competent  to  run  an  engine ;  and  when 
the  starting  point  has  been  reached,  no  time  is  lost  in  coupling  it  to  another  train,  and 
thus  it  is  kept  running  almost  continually  in  all  kinds  of  weather.  Compare  this 
treatment  to  the  former  and  the  change  must  become  apparent. 

G.  The  passenger  engines  are  sometimes  subjected  to  the  same  severe  treatment, 
but  generally  an  engine  is  placed  in  the  charge  of  only  two  engineers,  one  of  these 
running  the  engine  during  one  trip,  and  the  other  having  charge  of  it  during  the  next 
trip,  and  so  relieving  each  other  alternately. 

7.  Allowing  different  engineers  to  run  the  same  engine  has  this  advantage,  namely : 
that  only  competent  engineers  can  hold  their  positions,  because  after  a  competent 
engineer  has  once  shown  what  the  engine  can  do,  the  other  engineers  must  make  the 
engine  perform  a  like  amount  of  work  in  the  same  time,  or  give  good  reasons  for  not 
doing  so.     Here  then  we  perceive  that  no  incompetency  is  admissible. 

8.  The  passenger  trains  are  also  heavier  now  than  in  former  times,  and  the  trips 
longer.     Generally  speaking,  all   engines  are   now  required  to  do  more  work  than 
formerly.     Engines  placed  in  such  severe  service  must  naturally  be  strong,  powerful 
and  durable,  well  put  together,  bolt  holes  reamed,  bolts  turned  and  fitted,  and  driven 
in  tightly.     In  the  modern  locomotive  the  boiler  is  larger  than  in  former  practice,  the 
frames  and  cylinders  are  heavier,  and  generally  all  working  parts  are  made  stronger. 


M<H>Kl;\    l.ni-n.wnr/l'K    CONSTRUCTION.  3 

0.  Ill  the  appearance  and  outside  finish  of  the  engine  we  also  notice  a  decided 
change.  For  instance,  (lie  landscape  paintings  and  pictures  of  birds  and  horses  on  the 
side  ot'  the  tender,  have  of  late  disappeared,  and  the  tanks  are  plainly  painted  with 
good  paint,  and  well  varnished. 

This  is,  in  the  writer's  opinion,  as  it  should  be,  because  pictures  on  the  side  of  the 
tank  seem  to  him  to  be  out  of  place.  A  tender  is  made  for  the  purpose  of  carrying 
water  and  fuel,  and  is  ill  adapted  for  a  picture-gallery.  The  brass  finish  on  the  engine 
and  fancy  ornaments,  such  as  eagles,  etc.,  are  also  things  of  the  past,  because  these 
require  too  much  time  and  expense  to  keep  clean  and  in  good  condition. 

From  these  remarks  the  reader  must  not  conclude  that  in  former  times  the  engines 
had  a  better  and  more  pleasing  appearance.  This  is  not  the  case,  because  years  of 
experience  have  exposed  faulty  constructions  in  former  locomotives,  which  h;,ve  been 
corrected  and  otherwise  improved  in  the  modern  engine,  and  since  correct  construc- 
tion and  distribution  of  metal  must  always  improve  the  appearance  of  a  machine,  we 
conclude  that  our  American  locomotives  as  now  built  (although  by  no  means  perfect) 
possess  elegance  in  form,  compactness  in  the  arrangement  of  the  different  pieces  of 
mechanism,  and  gracefulness  in  movement. 

CLASSIFICATION   OF  LOCOMOTIVES. 

10.  We  may  divide  the  different  kinds  of  locomotives  into  two  distinct  classes;  in 
one  class  we  may  place  the  ordinary  passenger  and  freight  locomotive,  and  in  the  other 
the  switching  engine  and  other  locomotives  designed  for  some  special  service. 

At  present  we  will  consider  only  the  first  class,  namely,  the  passenger  anu  freight 
locomotives.  These  engines  are  again  divided  into  four  different  classes,  namely :  1st, 
the  eight-wheeled  engine;  2d,  the  Mogul  engine;  3d,  the  ten-wheeled  engine;  4th,  the 
consolidation  engine. 

An  eight-wheeled  engine,  sometimes  called  the  American  locomotive,  because  this 
design  was  first  brought  out  in  this  country  and  used  here  more  than  elsewhere,  is  an 
engine  that  has  four  driving  wheels  and  four  truck  wheels,  as  shown  in  Fig.  1.  On 
some  roads  the  eight-wheeled  engine  is  used  for  both  passenger  and  freight  service, 
1ml  generally  it  is  recognized  as  the  passenger  engine.  A  Mogul  engine  is  an  engine 
that  has  six  driving  wheels  and  two  truck  wheels,  as  shown  in  Fig.  2.  These  engines 
arc  used  principally  for  freight  service;  occasionally  they  are  used  for  passenger 
service,  but  generally  they  are  recognized  as  freight  engines. 

A  ten-wheeled  engine  is  one  that  has  six  driving  wheels  and  four  truck  wheels,  as 
shown  in  Fig.  3.  Ten-wheeled  engines  are  used  for  fast  freight  service,  for  hauling 
heavy  passenger  trains,  or  for  a  mixed  traffic. 

A  consolidation  engine  is  an  engine  that  has  eight  driving  wheels  and  two  truck 
wheels,  as  shown  in  Fig.  4.  These  engines  are  used  for  heavy  freight  service  on  roads 
having  steep  grades. 

Now,  notice  the  eight- wheeled  engine  and  the  Mogul  engine;  each  one  has  eight 
wheels.  In  the  passenger  engine  four  wheels  of  the  whole  number  are  driving  wheels, 
and  in  the  Mogul  engine  six  wheels  of  the  whole  number  are  driving  wheels. 

Again,  notice  the  ten-wl led  engine  and  the  consolidation  engine;  each  one  has 


Fifj.l 


EIG II T-  WHEELED  EX 'G INK 

Fig.  2 


Jliffid-Wlicel—Rase 

Total  Wheel- Base 


—  -Itiyltl- Wheel- llaxr 

—  Total  Whvvl-Itase — 
COKSOLIDAXIOX  ENGINE 


<-<>.\sri;n  n<>\. 


ten  wheels.     In  th<-  ten-wheeled  engine  six  wheels  of  I  lie  whole  number  are  driving 
wlieels,  ;IIK!  lii  the  consolidation  eight  wheels  of  the  whole  number  ai*e  driving  wheels. 

WHEEL  BASE. 

11.  The  rigid  wheel  base  of  any  engine  is  the  distance  from  the  center  of  rear  to 
the  center  of  the  front  driving  wheel,  plainly  shown  in  figures.     The  total  wheel  base 
of  any  engine  is  the  whole  distance  from  the  center  of  the  rear  driving  wheel  to  the 
center  of  the  front  truck  wheel,  also  plainly  shown  in  figures. 

DATA  REQUIRED. 

12.  Before  we  can  decide  what  type  of  a  locomotive  to  adopt,  and  before  we  can 
determine  the  dimensions  of  this  engine,  we  must  know  the  following  particulars: 
1st,  the  total  weight  of  the  train — that  is,  the  combined  weight  of  the  load  and  cars; 
2d,  the  speed  of  train;  3d,  the  grades  and  curves  of  the  road  on  which  the  engine 
is  to  run  ;  4th,  gauge  of  track — that  is,  the  exact  distance  between  the  rails;  5th,  the 
weight  on  the  drivers  that  the  rails  of  the  road  can  hear;   (5th,  kind  of  fuel  to  be  used  ; 
7th,  kind  and  height  of  couplings  of  cars;  8th,  limitations,  if  any,  in  width,  height, 
length,  etc.,  by  tunnels,  overhead  bridges,  turn-tables,  etc. 

For  the  sake  of  simplicity,  let  us  first  find  the  type  and  the  dimensions  of  a 
locomotive  cajiable  of  hauling  a  train  of  given  weight  on  a  straight  and  level  track, 
leaving  the  speed  and  all  other  particulars  out  of  the  question. 

TRAIN   RESISTANCE. 

13.  The  principal  resistance  which  a  locomotive  must  overcome  in  slowly  hauling 
a  I  rain  over  a  straight  and  level  road,  is  rolling  and  axle  friction.     Hence  the  resistance 
to  motion  of  a  train,  or  the  train  resistance,  is  simply  rolling  and  axle  friction  combined. 
But  it  must  be  remembered  that  when  a  train  is  to  run  fast,  or  against  strong  winds, 
other  forces  must  be  overcome. 

By  rolling  friction  is  meant  the  resistance  to  motion  that  takes  place  where  the 
circumference  of  the  car  wheel  comes  in  contact  with  the  rail.  Axle  friction  is  the 
resistance  to  motion  that  takes  place  between  the  axle  journal  and  its  bearing. 

14.  An  ordinary  train,  composed  of  cars  whose  wheels  are,  say,  from  28  inches  to 
'•'>-  inches  in  diameter,  and  having  journals,  say,  from  .'!  inches  to  3J  inches  in  diame- 
ter, will  require  a  force  of  7£  pounds  for  every  ton  of  2,000  pounds  to  move  it.     Thus, 
for  instance,  if  the  total  weight  of  the  cars  and  the  load  is  1,000  tons,  we  have  1,000 
x  74  =  7,500  pounds ;   this  means  that  a   train  of  1,000  tons   requires  a  force  of  7,500 
pounds  to  move  it,  or,  in  other  words,  it  requires  a  force  of  7,500  pounds  to  overcome 
the  combined  rolling  and  axle  friction. 

On  some  roads  it  may  require  only  <>  pounds  for  every  ton,  and  on  other  roads  it 
may  require  !)  pounds  to  move  a  ton  weight.  This  difference  is  caused  by  the  degree 
of  smoothness  and  irregularities  of  the  rails,  the  different  proportions  of  the  wheels 
and  journals,  the  kind  of  springs  under  the  cars,  the  kind  and  quantity  of  oil  used. 


Fig.  5 


6  MODERN  LOCOMOTIVE   COXSTRUCTION. 

and  other  minor  conditions.  We  believe  that  1\  pounds  per  ton  will  bo  suitable  for 
the  average  railroads,  and  this  figure  we  shall  hereafter  adopt  in  all  our  calculations 
in  which  speed  and  the  grade  is  not  taken  into  account. 

15.  The  amount  of  the  combined  rolling  and  axle  friction  of  a  train,  which  a  loco- 
motive must  overcome,  can  be  found  by  several  practical  methods.  For  instance: 
Assume  that  one  end  of  a  rope  is  attached  to  a  car,  and  the  other  end,  c,  of  rope  passed 
over  pulley,  b,  as  shown  in  Fig.  5.  The  bearings  of  this  pulley  are  supposed  to  be 

firmly  fastened  to  the  track,  and  the  height  and 
position  of  the  pulley  being  such  that  portion  a  // 
of  the  rope  will  be  parallel  to  the  rail;  then  a 
weight  fastened  to  the  end,  c,  of  the  rope,  and 
sufficiently  heavy  to  move  the  car,  and  no  more, 
will  be  the  force  in  pounds  necessary  to  move  it,  or, 
in  other  words,  this  weight  will  be  the  force  neces- 
sary to  overcome  the  combined  rolling  and  axle 
friction.  Hence,  if  the  weight  of  this  car  is  20  tons, 
we  may  expect  to  find  that  a  weight  from  120  to  180  pounds  will  move  it,  this  dif- 
ference of  weight  being  caused  by  the  conditions  of  the  rail,  etc.,  as  before  explained. 
Now,  the  mean  between  120  and  180  is  150  pounds  to  move  20  tons,  which  is  equiva- 
lent to  7£  pounds  per  ton. 

Again,  we  may  try  another  method.  Instead  of  placing  a  coiipling-bar  between 
the  tender  and  cars,  let  us  couple  these  by  an  instrument  capable  of  measuring  a  force. 
Such  an  instrument  is  called  a  dynamometer.  There  are  different  kinds  of  dyna- 
mometers, the  simplest  being  a  spring  balance,  sufficiently  strong  to  withstand  the 
pull,  and  yet  elastic  enough  to  indicate  correctly  the  force  in  pounds  exerted  by  the 
engine  in  pulling  the  train.  Although  the  spring  balance  is  not  always  the  best  instru- 
ment to  use  for  this  purpose,  and  is  adapted  only  for  moderate  forces,  we  draw  atten- 
tion to  it  because  its  action  is  familiar  to  the  reader,  and  probably  best  understood. 
Now  suppose  a  correct  spring  balance  is  placed  between  the  tender  and  a  train  whose 
weight  is  1,000  tons,  then,  as  soon  as  the  engine  commences  to  pull  and  move  the  train, 
our  spring  balance  will  show  a  force  from  6,000  to  9,000  pounds.  The  mean  between 
0,000  and  9,000  pounds  is  7,500  pounds,  which  is  again  equivalent  to  7£  pounds 
per  ton. 

We  may  also  determine  by  observation  the  force  necessary  to  move  the  train.  It 
has  been  found  that  railroad  cars,  with  wheels  and  axles  as  before  described,  will  begin 
to  roll  down  a  grade  when  it  is  as  steep  as  from  16  to  24  feet  per  mile.  Of  course,  this 
difference  is  caused  by  the  condition  of  the  track  and  other  considerations  before  men- 
tioned. 

Let  the  length  of  the  line  a  c,  Fig.  6,  represent  a  mile,  and  the  length  I  c  the  rise 
of  the  grade,  namely,  16  feet.  In  that  branch  of  science  called  mechanics  it  has  been 
proved  that  the  force  necessary  to  overcome  friction  is  as  much  smaller  than  the 
weight  *  of  the  cars  as  the  length  of  the  line  I  c  is  shorter  than  the  length  of  the  line  a  c. 

*  Instead  of  the  word  "weight,"  we  should  have  said  "pressure,"  because  the  weight  and  pressure  arc  equal 
only  on  a  level  track,  and  not  on  a  grade;  Imt  in  this  particular  case,  the  difference  being  so  small,  we  have,  for 
The'  sake  of  simplicity,  used  the  word  "weight."  How  to  find  this  difference  will  be  explained  hereafter. 


M<)l>l-:i:\    l.tti-tt.MoTII'K    CONSTRUCTION.  7 

Now,  the  line  n  <•  represents  one  mile,  or  .">,2SO  feet,  .-mil  (he  line  It  c  1G  feet;  dividing 
r>,2SO  ft 'ft   liy  It!  t'cft,  \vf  liavt-  '  "'.    =  :!:>(),  that  is,  the  line  a  c  is  330  times  longer 

than  tin-  line  l>  c,  hfiicf  the  weight  of  the  cars  will  be  330  times  greater  than  the  force 

''000 
necessary  to  overcome  friction.    Now,  there  are  12,000  pounds  in  a  ton,  hence  -    rr  =  6.06. 

oou 

This  nn'.nis  that  it  requires  (i  pounds  per  ton  to  move  the  train.     If  we  assume  the 
grade  to  lie  24  feet  in  a  mile  when  the  cars  begin  to  roll  down  the  grade,  then  the  line 

5280 

It  r,  Fig.  (i,  will  he  '24  feet  long,  and    ~       =  220,  that  is,  the  line  n  c  is  220  times  longer 

&*• 

''000 
than  the  line  \>  r,  therefore  ~^-r  —  J)  pounds  per  ton  to  move  the  train,  or  to  overcome 

rolling  and  axle  friction;  the  mean  between  6  and  9  pounds  is  7£  pounds,  as  before. 

From  this  we  may  establish 

a  rule  for  finding  the  force  nee-  Fiy.H 

essary  to  overcome  the  train 
resistance,  the  speed  not  being 
taken  into  consideration. 

RULE  1. — Multiply  the  weight  of  the  train  in  tons  (of  2,000  pounds)  by  7£;  the 
answer  will  be  (lie  force  in  pounds  necessary  to  overcome  the  train  resistance.  If  to 
this  we  add  th"  resistance  of  the  lender  and  its  load,  also  the  force  necessary  to  move 
the  engine  itself,  we  then  know  the  force  an  engine  must  exert  to  haul  the  total  load. 

For  all  practical  purposes  we  may  assume  that  7 A  pounds  per  ton  is  not  only  suffi- 
cient to  move  the  train,  but  also  includes  the  force  necessary  to  move  the  engine 
and  overcome  the  friction  of  its  machinery,  hence  no  separate  calculation  for  this  is 
necessary. 

The  resistance  of  the  tender  is  found  by  Rule  1 — that  is,  multiply  the  weight  in 
tons  of  the  tender  and  its  load  by  7J,  and  the  answer  will  be  the  force  in  pounds  re- 
quired t<>  overcome  this  resistance.  Or,  still  simpler,  add  the  weight  of  the  tender  and 
its  load  to  the  weight  of  the  train  and  multiply  the  sum  by  7i.  Thus, 

K\AMi'LE  1. — The  weight  of  a  train  is  1,200  tons,  and  the  weight  of  the  tender 
2(1  tons  ;  find  the  force  in  pounds  necessary  to  haul  this  train  ;  1,200  4-  20  =  1,220  tons, 
1,221)  x  7i  =  !),!")()  pounds,  hence  the  engine  must  be  capable  of  exerting  a  total  pulling 
force  of  !),!.")()  pounds. 

ADHESION. 

16.  The  effort  to  haul  a  train  which  a  locomotive  can  exert  is  limited  by  the 
adhesion  between  the  driving  wheels  and  the  rails.  This  adhesion  is  simply  friction 
between  the  driving  wheels  ami  rails  acting  so  as  to  prevent  slipping.  If,  for  instance, 
the  train  resistance  exceeds  the  adhesion,  the  driving  wheels  will  slip,  or,  in  other 
words,  turn  round  without  advancing. 

The  adhesion  depends  upon  the  weight  placed  on  the  drivers.  When  the  rails  are 
dry  and  in  comparatively  good  condition,  we  may  assume  that  the  adhesive  force  is 
equal  to  1  -of  the  weight  on  the  drivers.  Thus,  for  instance,  if  the  weight  on  the 
drivers  is  40,000  pounds,  the  adhesive  force  will  be  S,000  pounds.  This  adhesive  force 
enables  an  engine  to  pull  a  train,  and  must  not  be  less  than  the  train  resistance. 


g  MODERN  LOCOMOTIVE   CONSTRUCTION. 

When  the  rails  are  wet,  muddy,  or  greasy,  this  adhesive  force  will  be  considerably  less, 
and  snowy  or  frosty  weather  will  also  reduce  the  adhesion. 

In  the  following  calculations  we  shall  consider  the  track  to  be  in  good  condition, 
and  therefore  shall  assume  the  adhesion  to  be  equal  to  .'  of  the  weight  on  the  drivers. 
If  the  condition  of  the  track  is  not  known,  the  writer  believes  that  the  adoption  of  -,V  of 
the  weight  on  the  drivers  for  the  adhesion  will  not  lead  to  disappointment  as  often  as 
when  \  is  adopted. 

WEIGHT   ON   DRIVERS. — NUMBER   OF   DRIVING   WHEELS. 

17.  From  the  foregoing  remarks  we  have  learned  that  when  the  weight  of  the  train 
and  tender  is  known  we  can  find  the  train  resistance ;  also,  that  the  adhesion  must  at 
least  be  equal  to  the  train  resistance,  and  since  the  adhesion  is  equal  to  i  of  the  weight 
on  the  driving  wheels,  we  multiply  the  train  resistance  or  the  adhesion  by  5,  the  prod- 
uct will  be  the  total  weight  on  all  the  drivers. 

EXAMPLE  2. — In  Example  1  we  found  the  train  resistance  to  be  9,150  pounds ;  what 
must  be  the  total  weight  on  the  driving  wheels?  9,150  x  5  =  45,750  pounds,  hence  the 
total  weight  in  all  the  drivers  will  be  45,750  pounds. 

On  some  roads  heavy  rails  are  used,  on  other  roads  lighter  rails  are  adopted.  The 
heavy  rails  can,  of  course,  bear  a  greater  weight  on  the  drivers  than  the  lighter  rails, 
therefore,  before  we  can  find  the  number  of  drivers  under  an  engine,  we  must  know 
the  weight  that  the  rails  can  bear. 

18.  When  an  engine  is  running  on  light  rails — about  30  pounds  per  yard — we  may 
place  4,000  pounds  on  each  driver ;  and  when  an  engine  is  running  on  heavy  rails  we 
may  place  15,000  pounds  on  each  driver.     In  late  years  the  tendency  has  been  to  crowd 
all  the  weight  on  the  drivers  that  can  possibly  be  placed  on  them,  so  that  now  on  some 
roads  more  than  15,000  pounds  ai-e  placed  on  a  driver.     But  there  must  be  a  limit  to 
this  weight,  because  when  too  much  weight  is  placed  on  the  drivers,  either  the  tires, 
the  rails,  or  both,  will  be  injured.     The  exact  amount  of  weight  that  can  be  placed 
on  the  drivers  has  not  yet  been  satisfactorily  established,  but  we  believe  that  the  fore- 
going figures,  namely,  4,000  to  15,000  pounds  on  each  driver,  according  to  size  of  rail, 
may  be  safely  adopted. 

From  these  remarks  it  must  be  evident  that  before  we  can  decide  which  of  these 
two  figures  we  can  use,  or  what  amount  of  weight  between  these  two  limits  we  may 
adopt,  we  must  know  the  material  of  which  the  rails  are  made,  and  the  weight  of  rail 
per  yard,  that  is,  their  form  and  size.  Of  course,  we  are  now  alluding  only  to  rails  for 
ordinary  passenger  and  freight  engines  on  roads  of  3  feet  gauge,  or  other  roads  up  to 
4'  8£"  gauge,  and  we  do  not  include  the  rails  for  mining  engines,  plantation  engines, 
or  wooden  rails. 

Another  important  fact  that  we  must  not  overlook  is  the  weight  the  bridges  can 
bear,  because  the  rails  may  be  suitable  for  a  heavy  load,  and  the  bridges  may  not  be  so. 

19.  If,  then,  we  know  the  weight  that  can  be  safely  placed  on  each  driver,  we  can 
find  the  number  of  drivers  to  be  placed  under  an  engine  by : 

RULE  2. — Divide  the  weight  that  must  be  placed  on  all  the  drivers  by  the  weight 
that  can  be  safely  placed  on  one  driver,  and  the  quotient  will  be  the  number  of  driving 
wheels  required. 


K  .'!. — The  .greatest  weight  <>n  each  driver  that  the  rails  of  a  given  road  can 
hear  is  1(1,1100  pounds,  and  the  weight  necessary  on  all  the  drivers  to  haul  the  train  is 
40,000;  how  many  driving  wheels  must  be  placed  under  the  engine!  According  to 

1 1  |,  w  w  \ 

the  rule  we  have         -  =  4,  hence  the  number  of  drivers  will  be  four.     If  the  necessary 
1 0(  M  K) 

weight  on  all  the  drivers  had  been  (iO,000  pounds,  we  then  would  have  to  place  six 
drivers  under  the  engine  so  as  not  to  exceed  10,000  pounds  on  each. 


DIAMETERS   OF   DRIVING   WHEELS. 

20.  The  diameter  of  the  driving  wheels  under  an  engine  will,  to  a  great  extent, 
depend  upon  tin-  speed  of  the  locomotive.  Driving  wheels  of  large  diameter  are  neces- 
sary for  fast  speeds;  and,  on  the  other  hand,  driving  wheels  for  heavy  freight  engines 
must  necessarily  be  comparatively  small  in  diameter.  There  are  several  causes  which 


Fig.  7 

will  place  a  limit  to  the  diameter  of  a  driving  wheel  in  either  direction.  We  will  name 
two:  The  diameter  must  not  be  too  large,  because,  if  it  is,  the  engines  will  stand  too 
high.  The  diameter  must  not  be  too  small,  because,  if  it  is,  difficulty  will  be  experi- 
enced in  getting  steam  out  of  the  cylinder  on  account  of  the  high  piston  speed  which 
may  be  necessary  for  the  required  speed  of  train.  Between  these  two  limits  no  exact 
rule  for  finding  the  diameter  of  a  driving  wheel  can  be  given.  The  following  tables 
will  greatly  aid  us  in  determining  the  diameters  of  these  wheels.  These  tables  show 
the  diameters  of  driving  wheels  for  the  different  classes  of  engines,  such  as  are  gener- 
ally adopted  by  builders  and  master  mechanics,  and  giving  good  satisfaction. 

In  these  tables  we  see  that,  for  an  eight-wheeled  engine  with  cylinders  10"  in 
diameter  and  20"  stroke,  we  may  use  driving  wheels  45"  diameter,  or  larger,  up  to  51" 
diameter;  or,  if  it  is  an  eight-wheeled  engine  with  17"x24"  cylinders,  we  may  adopt 
driving  wheels  (i()"  diameter,  or  larger,  up  to  (50".  Of  course,  these  limits  of  driving 
wheels  for  the  different  classes  of  engines  are  not  absolute.  We  may  change,  and, 
indeed,  may  be  compelled  to  change,  these  diameters  to  suit  some  particular  service. 

But  it  must  be  remembered  that  when  the  number  of  revolutions  of  the  driving 
wheel  per  mile  are  given,  then  the  diameter  of  the  driver  is  not  a  matter  of  choice,  but 
must  be  found  accurately  l,v  calculation,  which  is  an  easy  matter.  Thus,  for  instance, 
the  number  of  revolutions  of  the  driving  wheel  per  mile  is  :!:!<>:  what  must  be  the 
diameter  of  the  wheel  .'  One  mile  is  equal  to  5,2S()  feet;  then  dividing  5,280  by  the 

5*  'SO 
number  of  revolutions,   namely,   :>:!(>,   we  have  -     -  =  15.71.      This    quotient  is  the 


number  of  feet  in  the  circumference  of  the  wheel.  Now,  if  we  refer  to  a  table  of 
circumferences,  we  find  that  the  diameter  of  a  circle,  whose  circumference  is  15.71 
feet,  is  equal  to  5  feet.  If  such  a  table  is  not  at  hand,  then  divide  the  15.71  feet  by 


10 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


3.1416,  because  the  circumference  is  always  3.141  G  times  greater  than  the  diameter,  and 
the  quotient  in  this  case  will  be  5  feet,  which  is  the  diameter  of  the  wheel. 


TABLES    SHOWING   THE    DIAMETERS   OP    DRIVING   WHEELS   AS    GENERALLY   ADOPTED    FOR    DIF- 
FERENT  CLASSES   AND   SIZES   OF   LOCOMOTIVES. 

ALL  DIMENSIONS  IN  INCHES. 


TABLE  1. 
EIGHT-WHEELED  ENGINES. 

TABLE  2. 
MOGUL  ENGINES. 

TABLE  3. 
TEN-WHEELED  ENGINES. 

TABLE  4. 
CONSOLIDATION  ENGINES. 

Cylinder?. 
Diameter.  Stroke. 

Driving  Wheels. 
Diameter. 

Cylinders. 
Diameter  Stroke. 

Driving  \Vh<-rle. 
Diameter. 

Cylinders. 
Diameter.  Stroke. 

Driving  Wheels. 
Di:i  meter. 

Cylinders. 
Diumeter.  Siroke. 

Driving  Wheels. 
Diameter. 

Column  1. 

Column  a. 

Column  ]. 

Column  2. 

Column  1. 

Column  2. 

Column  1. 

Column  2. 

10  x  20 
11  x  22 
12  x  22 
13  x  22 
14  x  24 
15  x  24 
16  x  24 
17  x  24 
18  x  24 

45  to  51 
45  to  51 
48  to  54 
49  to  57 
55  to  61 
55  to  66 
58  to  66 
60  to  66 
61  to  66 

11  x  16 
12  x  18 
13  x  18 
14  x  20 
15  x  22 
16  x  24 
17  x  24 
18  x  24 
19  x  24 

35  to  40 
36  to  41 
37  to  42 
39  to  43 
42  to  47 
45  to  51 
49  to  54 
51  to  56 
54  to  60 

For  Narrow  Gauge. 

t       3'  0"  to  3'  0" 

12  x   18 
13  x  18 
14  x  20 
15  x  22 
16  x  24 
17  x  24 
18  x  24 
19  x  24 

39  to  43 
41  to  45 
43  to  47 
45  to  50 
48  to  54 
51  to  56 
51  to  56 
54  to  60 

]  14   x  16 
(15  x  18 

:;ii  to  38 
36  to  38 

For  4'  8i"  Gauge. 
)  20  x  24     48  to  50 
\  22  x  24     50  to  52 

The  reason  why  5,280  is  divided  by  the  number  of  revolutions  per  mile  is  simple, 
and  yet  it  is  not  so  generally  understood  among  mechanics  as  we  might  expect,  there- 
fore the  following  explanation  is  offered : 

Let  the  line  a  &,  Fig.  7,  represent  one  mile — that  is,  5,280  feet — and  c  the  center  of 
the  wheel  when  it  stands  at  the  end,  a,  of  the  line  a  ft,  as  shown  in  the  figure.  Care- 
fully rolling  this  wheel  without  slipping  along  the  line  a  b  until  it  has  completed  one 
revolution,  or  made  one  complete  turn,  the  number  of  feet  that  it  has  traveled  along 
the  line  a  b  is  equal  to  the  number  of  feet  in  its  circumference.  Now,  if  this  wheel  is 
5  feet  in  diameter,  its  circumference  will  be  15.71  feet,  nearly,  and  the  distance  from  the 
center  c  to  the  center  co,  which  is  equal  to  the  distance  that  it  has  traveled  along  the  line 
a  b,  will  be  15.71  feet.  Again,  if  we  continue  rolling  this  wheel  along  the  line  a  b  until  it 
has  made  another  complete  turn,  then  its  center  will  be  at  cs,  and  the  distance  that  the 
wheel  has  traveled  from  the  point  a  along  the  line  a  b  will  be  equal  to  the  distance 
between  the  first  center  c  and  c* — that  is,  31.42  feet — which  is  obtained  by  multiplying 
15.71  x  2 ;  and  so  on  for  every  revolution  it  will  travel  15.71  feet  further ;  and  therefore, 
if  we  divide  the  length  of  the  line  a  b,  or  one  mile,  by  the  circumference  of  a  wheel, 
which  is,  in  this  case,  15.71  feet,  we  will  know  the  number  of  revolutions  that  it  must 
make  to  travel  from  a  to  b.  And  conversely,  if  we  know  the  number  of  revolu- 
tions per  mile,  we  divide  the  number  of  feet  in  a  mile,  or,  as  in  our  example,  the 
length  of  the  line  a  b  by  the  number  of  revolutions,  and  the  quotient  will  lie  the 
circumference  of  the  wheel,  aud  dividing  this  circumference  by  3.1410,  will  give  its 
diameter. 


Tli.UTIVF.    POWER. —  DIAMETER    OF    CYLINDERS. 

•Jl.  \\'c  now  come  tn  the  consideration  of  the  size  of  cylinders  necessary  to  turn 
the  driving  wheels.  Neglecting  the  friction  of  the  niiicliinciy,  we  may  say  that  the 
cylinders  with  a  given  steam  pressure  must  be  large  enough  to  almost  slip  the  wheels, 
that  is,  to  turn  the  wheels  without  advancing  on  the  rails,  when  the  engine  is  attached 
to  the  heaviest  train  that  it  was  designed  to  haul.  Or,  in  other  words,  if  a  certain 
amount  of  weight  is  placed  on  the  drivers  to  haul  a  given  train,  we  must  design  our 
cylinders  so  that  a  sullicieut  power  can  be  obtained  to  turn  the  wheels,  and  not  more 
and  not  less,  when  the  engine  is  attached  to  this  train.  This  power  which  is  necessary 

to  turn  the  driving  \vl Is  under  the  above  conditions  is  called  the  "  tractive  power"of 

a  locomotive.  If  the  cylinders  are  too  small  in  proportion  to  the  weight  placed  on  the 
drivers,  th.'n  the  engine  cannot  haul  the  train  that  it  was  intended  it  should  do  with 
the  correct  weight  on  the  drivers.  If,  on  the  other  hand,  the  cylinders  are  too  large 
in  proportion  to  the  weight  placed  on  the  drivers,  then  the  engine  cannot  employ  all 
its  tractive  power.  In  these  cases,  there  will  be  either  a  waste  of  material  or  steam. 
Here,  then,  we  see  that  in  a  correctly  designed  engine  there  is  a  fixed  relation  or 
proportion  between  its  tractive  power  and  the  weight  placed  upon  the  drivers.  The 
tractive  power  is  not  only  dependent  upon  the  diameter  of  the  cylinders,  but  also 
upon  the  diameters  of  the  drivers,  the  length  of  stroke,  and  the  mean  effective  steam 
pressure  per  square  inch  of  piston. 

'2'2.  In  Fig.  S  we  have  represented  a  pair  of  cylinders  and  one  of  the  front  pair  of 
driving  wheels  of  an  eight-wheeled  engine,  such  as  shown  in  Fig.  1.  One  of  the  cylin- 
ders in  Fig.  8  is  connected  to  the  driving  wheel;  the  other  cylinder  is  connected  to  a 
crank  fastened  to  the  same  axle,  and  not  connected  to  a  driving  wheel,  because  we  have 
assumed  that  there  is  only  one  driving  wheel  on  the  axle.  Let  us  also  assume  that  the 
cylinders  in  Fig.  S,  with  frames,  valve  gear,  and  all  necessary  mechanism,  are  firmly 
fastened  to  blocks  or  a  foundation,  so  that  this  figure  represents  a  complete  stationary 
engine.  The  driving  wheel  is  not  to  touch  the  track,  but  the  whole  engine  is  set  high 
enough  so  that  a  rope  can  be  fastened  to  the  lower  part  of  the  driving  wheel,  in  such  a 
manner  that  when  the  other  end  of  the  rope  is  attached  to  a  train  this  rope  will  be 
parallel  to  the  track,  as  shown.  When  this  engine  is  set  in  motion  in  a  direction  as 
shown  by  the  arrow,  it  will  haul  the  train  towards  the  engine.  Now,  the  power  that 
this  engine  exerts  in  doing  this  is  precisely  the  same  as  the  tractive  power  of  a  locomo- 
tive designed  to  do  the  same  amount  of  work.  Should  the  total  weight  of  this  train 
be  1,000  tons,  then,  according  to  what  has  been  said  before,  it  will  take  7,500  pounds 
to  move  it,  and  therefore  the  stress  or  the  pull  on  the  rope  will  be  7,~>00  pounds. 

If,  instead  of  fastening  this  rope  to  the  train,  we  pass  it  over  a  pulley,  »,  and  attach 
a  weight,  tr,  to  it  weighing  7,f>00  pounds,  as  shown  by  the  dotted  lines  in  Fig.  8,  then 
the  stress  or  the  pull  on  the  rope  will  not  be  changed,  but  will  still  remain  as  before, 
namely,  7, .">!)()  pounds,  and  therefore  we  conclude  that  the  power  necessary  to  move  the 
train  is  exactly  the  same  as  the  power  necessary  to  hoist  a  given  weight. 

For  the  sake  of  clearness  and  simplicity,  when  calculating  the  tractive  power  of  a 
locomotive,  we  shall  hereafter  always  assume  that  the  train  resistance  is  represented 
by  a  weight,  u;  fastened  by  the  means  of  a  rope  to  t»tc  driving  wheel,  as  shown  in 


12 


MODERN  LOCOMOTIVE   COXSTRVCTION. 


Fig.  8,  and  that  the  cylinders  must  be  made  large  enough  so  as  to  be  capable  of 
lifting  this  weight. 

But  the  reader  may  say  that  a  locomotive  has  more  work  to  do  than  the  stationary 
engine  here  represented,  because  the  locomotive  must  move  its  own  weight,  which  the 
stationary  engine  does  not  have  to  do.  This  is  true,  but  it  must  be  remembered  that  we 
are  allowing  7|  pounds  for  every  ton  of  the  weight  of  train  that  is  to  be  moved,  and,  as 
we  have  stated  before,  this  may  be  considered — and  we  do  consider  it  so — as  not  only 
sufficient  to  move  the  train,  but  also  sufficient  to  move  the  weight  of  the  locomotive 
and  overcome  the  friction  of  its  mechanism.  Yet  we  must  again  call  the  attention  of 
the  reader  to  the  fact,  that  any  particular  or  given  speed  is  not  yet  taken  into  con- 
sideration; we  are  simply  proportioning  an  engine  capable  of  moving  a  train  vcry 
slowly. 

23.  The  foregoing  being  thoroughly  understood,  the  solution  of  the  following 
example  will  not  be  difficult : 

EXAMPLE  4. — Find  the  diameters  of  the  cylinders  for  an  eight-wheeled  locomotive, 
whose  total  weight  on  drivers  is  20,000  pounds,  the  diameter  of  the  driving  wheels, 
45  inches ;  the  stroke,  20  inches ;  and  the  mean  effective  steam  pressure  90  pounds  per 


square  inch  of  piston  area.  (The  writer  believes  that  for  the  mean  effective  steam 
pressure  90  pounds  per  square  inch  is  a  good  average,  and  this  will  always  be  adopted 
unless  otherwise  stated.) 

From  what  has  been  said  before,  we  know  that  the  total  adhesive  force  will  be 
-f  of  the  weight  placed  on  the  drivers ;  hence  the  total  adhesive  force  will  be  £  of 
20,000,  which  is  equal  to  4,000  pounds.  We  have  also  seen  that  the  adhesion  is 
equal  to  the  train  resistance;  hence  the  weight  ^v,  in  Fig.  8,  which  represents  the 
train  resistance,  must  weigh  4,000  pounds.  Now,  all  we  have  to  do  is  to  find  the 
diameters  of  the  two  cylinders,  as  shown  in  Fig.  8,  capable  of  lifting  this  weight  of 
4,000  pounds;  and  so  for  all  locomotives  when  the  total  weight  on  the  drivers  is 
known,  no  matter  how  many  driving  wheels  are  to  be  placed  under  an  engine,  we 
always  assume  that  the  train  resistance  is  represented  by  the  weight  w;  that  all  this 
weight,  or  train  resistance,  is  applied  to  only  one  driving  wheel;  and  that  the  two 


13 

cylinders  must  In-  made  largo  enough,  so  that  their  combined  effort  mil  be  sufficient  for 
lifting  this  one  weight  it;  which  \vc  assume  to  be  4  of  the  total  weight  placed  on  all  the 
drivers. 

In  our  example,  as  we  have  already  seen,  this  weight  w  is  equal  to  4,000  pounds, 
and  the  diameters  of  the  driving  wheels  45  inches  each.  When  the  wheel  has  made 
one  revolution,  the  weight  tr  will  then  have  been  raised  through  a  distance  equal 
to  the  circumference  of  the  wheel,  and  this  circumference  is  141.37  inches,  or 
11.7S1  feet. 

lu  raising  this  weight,  a  certain  amount  of  energy  must  be  expended;  and,  to 
know  exactly  how  much  has  been  expended,  we  must  compare  it  to  some  standard  or 
unit  of  energy. 

The  amount  of  work  required  to  raise  or  lift  one  pound  one  foot  high  is  equal 
to  a  unit  <>f  energy  or  foot-pound;  hence,  if  two  pounds  are  lifted  one  foot  high,  two 
units  of  energy  have  been  expended,  or,  if  five  pounds  are  lifted  OIK*  foot  high,  five 
units  of  energy  have  been  expended;  and,  if  the  five  pounds  are  raised  five  feet  high, 
then  2.">  units  of  energy  have  been  expended,  because,  to  raise  the  five  pounds  through 
the  lirst  foot,  five  units  of  energy  will  be  required;  to  raise  them  through  the  second 
foot  another  five  units  will  be  required;  the  same  for  the  third,  and  so  on  up  to  the 
fifth,  making  a  total  of  23  units  of  energy,  or  foot-pounds. 

In  our  example  a  weight  of  4,000  pounds  must  be  raised  11.781  feet  high.  To 
raise  this  weight  through  the  first  foot,  4,000  units  of  energy  or  foot-pounds  will  be 
required;  and  the  same  amount  of  energy  will  be  required  to  raise  it  through  the 
second  foot,  and  again  the  same  through  the  third  foot,  and  so  on  until  the  height  of 
11.7S1  feet  has  been  reached ;  therefore,  the  total  number  of  units  of  energy,  or  foot- 
pounds, that  must  be  expended  to  raise  this  weight  through  a  distance  of  11.781  feet 
is  4,000  x  11.781  =  47,124  foot-pounds.  In  a  similar  wray,  for  all  engines,  we  multiply 
the  weight,  it;  in  pounds,  which  represents  the  adhesion,  by  the  circumference  of  the 
wheel  in  feet,  and  the  product  will  be  the  number  of  foot-pounds  or  units  of  energy 
that  must  be  expended  during  the  time  the  wheel  makes  one  revolution. 

Hut  the  energy  necessary  to  raise  this  weight  is  derived  from  the  steam  pressure 
in  the  cylinder,  and  since  the  mean  effective  steam  pressure  prr  square  inch  of  piston 
is  already  given — namely,  !M)  pounds — it  only  remains  to  make  the  cylinder  of  such  a 
diameter  that  we  can  obtain  47,124  units  of  energy  for  every  turn  of  the  wheel  with  a 
weight  attached,  as  shown  in  Fig.  8. 

But  now  notice  the  fact  that  during  the  lime  the  wheel  makes  one  turn,  raising 
the  weight  11.781  feet  high,  the  piston  travels  1  hrongli  a  distance  equal  to  twice  the 
length  of  the  stroke;  the  stroke  being  20  inches,  the  piston  travels  through  a  distance 
of  40  inches,  or  I!..'!,';  feet.  During  the  time  that  the  piston  travels  through  a  distance 
of  :!.•">:!  feet,  47,124  units  of  energy  or  foot-pounds  must  be  expended,  and  therefore, 

47124 

dividing  47,124  by  .'!.)!:!  feet,  we  have   — —  =  14,151  pounds.     This  last  answer  simply 

»).»>•> 

means  that  to  raise  the  4,000  pounds  weight  through  a  distance  of  11.781  feet,  the 
weight  being  attached  to  the  wheel,  as  shown,  will  require  as  many  units  of  energy 
as  to  raise  a  weight  of  14,1.">1  pounds  .'!.:!.'!  feet  high,  the  weight  being  attached  directly 
to  the  end  of  the  piston  rod,  as  shown  in  Fig.  9. 


14 


MODEltX   LOCOMOTIVE   CONSTRUCTION. 


Now,  it  will  be  readily  understood  that  the  mean  effective  steam  pressure  on  each 
square  inch  of  piston  will  lift  a  portion  of  this  weight  of  14,lf>l  pounds,  and  the 
amount  that  the  pressure  per  square  inch  of  piston  will  lift  is  90  pounds;  licncc,  divid- 
ing 14,151  pounds  by  90  pounds,  we  have  =  157.2  square  inches.  This  means 

t/\/ 

that  the  total  piston  area  must  be  157.2  square  inches.     But  we  have  two  cylindei's ; 

157  2 

therefore        —  =  78.6  square  inches  in  the  area  of  one  piston ;  and  a  piston  having  an 

area  of  78.6  square  inches,  must  be  10  inches  diameter.  Hence  a  locomotive  having 
four  driving  wheels,  with  20,000  pounds  placed  upon  them,  the  driving  wheels  being  4~> 

inches  in  diameter,  and  a  mean  effective  steam 
pressure  of  90  pounds  per  square  inch,  will 
require  cylinders  10  inches  in  diameter  and  20 
inches  stroke. 

Here  we  have  calculated  the  diameters  of 
the  cylinders  suitable  for  a  given  weight  placed 
on  the  drivers.  We  may  reverse  the  order  of  1  his 
calculation,  and  find  the  necessary  weight  that 
must  be  placed  on  the  drivers,  when  the  dimen- 
sions of  cylinders  and  diameters  of  driving 
wheels  are  given. 

EXAMPLE  5. — The  diameter  of  each  cylinder 

is  10  inches;  stroke,  20  inches ;  diameter  of  driving  wheels,  45  inches;  mean  effective 
steam  pressure,  90  pounds  per  square  inch  of  piston.  What  is  the  tractive  power  of 
such  an  engine  ?  And  how  much  weight  must  be  placed  on  the  drivers  I 

The  area  of  a  piston  10  inches  in  diameter  is  78.54  square  inches.  Multiplying  the 
area  of  the  piston  by  the  steam  pressure  per  square  inch,  we  have  78.54  x  90  =  7068.6 
pounds  total  steam  pressure  on  one  piston ;  but  there  are  two  pistons,  hence  7068.6 
x  2  =  14137.2  pounds,  which  is  the  total  steam  pressure  on  both  pistons.  The  stroke 
is  20  inches,  and  during  the  time  that  the  wheel  makes  one  turn,  the  piston  has 
traveled  through  twice  the  length  of  the  stroke ;  hence  20  x  2  =  40  inches,  or  3.33  feet. 
Multiplying  the  total  steam  pressure  on  the  pistons  by  3.33  feet,  we  have  14137.2  x 
3.33  =  47076.876  foot-pounds,  or  units  of  energy  the  cylinders  can  exert  during  one 
revolution  of  the  wheel.  The  driving  wheels  are  45  inches  in  diameter ;  hence  the  cir- 
cumference of  each  wheel  will  be  141.37  inches,  or  11.78  feet. 

Dividing  the  units  of  energy  the  cylinders  are  capable  of  exerting  by  the  circum- 
ference of  the  wheel,  we  have  -  -  —  3,996  *  pounds.  The  tractive  power  of  the 

11.78 

engine  is,  therefore,  capable  of  lifting  a  weight  of  3,996  pounds  attached  to  the  driving 
wheel,  as  shown  in  Fig.  8;  or,  in  other  words,  the  tractive  power  of  this  engine  is  capa- 
ble of  overcoming  a  train  resistance  of  3,996  pounds.  In  a  similar  manner,  the  tractive 
power  of  any  engine  may  be  found,  namely,  by  multiplying  together  twice  the  area  in 
square  inches  of  one  piston,  the  mean  effective  steam  pressure  per  square  inch,  and 

*  Tliis  answer  would  have  been  4,000,  instead  of  3,996,  if  the  decimal  fraction  iu  the  3.33  feet  (obtained  by 
multiplying  the  stroke  by  2)  had  been  exact. 


rn\sri;n-no.\.  15 

(win-  tlit-  length  of  tin-  stroke  in  feet;  then  dividing  this  product  by  the  circumference 
in  feet  of  the  wheel,  the  quotient  will  he  the  tractive  power  of  the  engine. 

This  rule  can  be  greatly  simplified,  as  we  shall  presently  show.  The  tractive! 
power  and  the  adhesion  are  represented  by  the  same  number  of  pounds;  therefore 
multiplying  the  tractive  power  by  .">,  we  have  :!,9!K>  x  :>=1!>,!)HO  pounds,  which  is  the 
weight  that  must  be  placed  on  the  drivers  of  this  particular  engine.  Ouv  answer,  then, 
to  Kxample  5  is:  Tractive  power,  3,996  pounds;  weight  on  drivers,  19,980  pounds. 

WEIGHT  OF  ENGINES. 

'24.  In  Kxample  4  it  has  been  shown  how  to  find  the  diameters  of  the  cylinders  when 
the  weight  on  the  drivers  and  the  diameters  of  the  same  are  known;  and  in  Example 
.">  it  has  been  shown  how  to  find  the  weight  on  the  drivers  when  the  dimensions  of  the 
cylinders  and  diameters  of  driving  wheels  are  known.  From  the  reasoning  connected 
with  these  examples,  we  conclude  that  the  tractive  power  should  not  be  more  or  less 
than  the  adhesion — a  fact  which  we  have  stated  before.  We  may  also  reasonably  con- 
clude that,  when  the  dimensions  of  the  cylinders  and  the  diameters  of  the  driving 
wheels  of  any  engine  are  given,  and  if  we  assume  that,  in  all  cases,  the  mean  effective 
steam  pressure  per  square  inch  of  piston  is  90  pounds,  we  may  at  once  arrange,  for 
future  use  and  reference,  a  table  for  each  class  of  engine,  showing  the  tractive  power 
of  each,  the  necessary  weight  on  the  drivers,  and  the  number  of  tons  of  2,000  pounds 
that  each  engine  can  haul  on  a  straight  and  level  track.  Indeed,  we  may  extend  these 
tables  so  that  the  weight  on  the  track,  and  consequently  the  total  weight  of  each 
engine,  will  at  once  be  seen. 

With  these  objects  in  view,  the  following  tables  have  been  prepared.  In  these 
tables,  columns  1  and  '2  are  exactly  the  same  as  those  given  in  tables  1,  2,  3  and  4.  In 
column  3  of  all  the  following  tables  the  adhesion  is  given  and,  since  the  adhesion  and 
tractive  power  are  expressed  by  the  same  number  of  pounds,  these  figures  are  obtained 
by  finding  the  tractive  power  of  each  engine,  and  for  this  purpose  the  small  diameters 
of  driving  wheels  given  in  column  2  are  always  used.  The  weight  on  drivers  is  shown 
in  column  4,  which  is  obtained  by  multiplying  the  adhesion  by  5  for  all  classes  of 
engines.  Column  5  gives  the  weights  on  the  trucks ;  the  calculations  for  these  weights  are 
based  upon  observations.  Thus,  it  has  been  noticed  that  the  weight  on  the  truck  for  an 
eight-wheeled  engine  is  about  one-half  of  that  placed  on  the  drivers ;  hence,  multiplying 
the  weight  placed  on  drivers  by  the  decimal  ..">,  the  weight  on  the  truck  will  be  known. 

For  Mogul  engines,  we  multiply  the  total  weight  on  drivers  by  the  decimal  .2,  and 
the  product  will  be  the  weight  on  the  truck. 

For  ten-wheeled  engines,  the  total  weight  on  the  drivers  multiplied  by  the  decimal 
..'!2  will  be  equal  to  the  weight  on  the  truck. 

And  lastly,  for  consolidation  engines,  the  total  weight  on  drivers  multiplied  by 
the  decimal  .!<>  will  determine  the  weight  on  the  truck. 

For  instance,  to  find  the  weight  on  the  truck  for  an  eight-wheeled  engine  with 
cylinders  17"x  24",  we  multiply  the  total  weight  on  drivers  for  this  engine,  given  in 
column  4,  by..">;  hence,  we  have  r>2,020  x.f>  =  2(i,010  pounds,  which  is  the  weight  on 
truck. 


1(3  MODEM  LOCOMOTIVE   CONSTRUCTION. 

For  a  17"  x  24"  Mogul  engine,  we  have  63,697  x  .12=1:2,7:!!)  pounds  =  weight  on 
truck. 

For  an  18"  x  24"  ten- wheeled  engine,  we  have  68,611  x  .32  =  21,95.")  pounds  = 
weight  on  truck. 

And  for  a  20"  x  24"  consolidation  engine,  we  have  90,000  x  .16  =  14,400  pounds  = 
weight  on  truck. 

In  column  6,  the  total  weight  of  each  engine  is  given,  which  is  obtained  by  adding 
the  weight  on  the  drivers  to  the  weight  on  the  truck.  Dividing  the  adhesion  given  in 
column  3  by  7J,  will  give  us  the  number  of  tons  of  2,000  pounds  that  the  engine  is 
capable  of  hauling  on  a  straight  and  level  track;  these  figures  arc  given  in  column  7. 

The  weight  of  engines  given  in  these  tables  will  be  found  to  agree  generally  very 
closely  with  the  actual  weights  of  locomotives  recently  built,  although  it  must  not  be 
expected  that  these  weights  will  agree  in  every  case  with  the  actual  weights,  because 
different  builders  do  not  build  their  engines  alike. 

The  given  weights  on  the  trucks  for  Mogul  or  consolidation  engines  may  differ 
considerably  from  the  actual  weights,  yet  this  should  not  be  a  matter  of  surprise, 
because  the  weight  on  a  truck  for  either  of  these  engines  can  be  changed  without 
changing  the  total  weight  of  engine,  and,  indeed,  often  the  pieces  of  mechanism  con- 
necting the  truck  to  the  engines  are  so  arranged  that  a  heavy  or  light  weight  can  be 
thrown  on  the  truck  while  the  engine  is  standing  on  the  track. 

Yet  the  figures  in  these  tables,  indicating  the  weights  on  the  trucks,  can  be  safely 
taken  as  guides  in  constructing  and  proportioning  an  engine. 

The  actual  weight  on  trucks  for  eight-wheeled  or  ten-wheeled  engines  will  not 
differ  much  from  those  given  in  the  tables,  because  these  weights  depend  greatly  on 
the  difference  between  the  total  and  rigid  wheel  bases,  and  these  are  not  often  changed 
by  the  different  builders.  The  ratio  of  the  rigid  and  total  wheel  buses  is  generally  the 
same  in  all  eight-wheeled  engines,  and  the  same  may  be  said  of  ten-wheeled  engines. 

It  has  already  been  stated  that  the  rule  (as  before  given)  for  finding  the  tractive 
power  of  an  engine  can  be  greatly  simplified.  To  explain  this,  we  will  again  state  the 
former  rule,  but,  instead  of  writing  it  in  the  ordinary  language,  we  will  employ  a  num- 
ber of  simple  arithmetical  signs.  This  will  enable  us  to  bring  the  whole  mode  of 
operation  under  the  eye,  and  follow  it  without  taxing  our  memory ;  hence  this  rule  will 
read  as  the  following : 

EULE  A. 

Area  of  piston  in  so.  in.  x  mean  effective  steam  pressure  per  sc[.  in.  x  stroke  in  ft.  x  L'  •  2 

—  =  tractive  power. 
Circumference  of  driving  wheel  in  feet 

For  the  sake  of  distinction  we  have  called  this  Rule  A.  Now,  remembering  that 
the  area  of  a  piston  is  found  by  multiplying  the  square  of  its  diameter  by  .7854,  and 
also  that  the  circumference  of  a  wheel  is  found  by  multiplying  its  diameter  by  3.1416, 
we  can  put  in  the  place  of  "area  of]>i*ti»i  hi  xi/ittin1  inches,"  in  Rule  A,  the  method  of 
finding  this  area,  namely,  square  of  diameter  in  inches  x  .7854;  and,  in  the  place  of 
"cireinnji'rciifc  of  irlu'rl  in  fret,"  we  may  put  diameter  of  wheel  in  feet  x  3.1416;  conse- 
quently, the  wording  of  the  Rule  A  will  be  changed,  and  read  like  Rule  B : 


<  <>\sri;r<'n<>\. 


17 


RULE  B. 

Sq.  of  diam.  of  piston  in  in.  x  .7854  x  mean  effective  steam  pressure  |>er  sq.  in.  x  stroke  in  ft.  x  2  x  2 
Diameter  of  driving  wheel  in  feet  x  3.141(1 


=  tractive  power. 


If,  now,  we  multiply  the  decimal  .7854  by  2,  and  again  by  2  (figures  which  are 
found  above  the  line  in  Rule  B),  we  have  a  product  of  3.1416.  Below  the  line,  in  Rule 
B,  we  find  the  same  figure — that  is,  3.1416 ;  hence  we  may  cancel  all  these  figures,  or, 
in  other  words,  we  may  throw  out  of  the  Rule  B  the  decimal  .7854  and  the  figures  2x2, 
all  found  above  the  line,  and  the  figure  3.1416,  which  is  found  below  the  line.  Doing 
so,  the  wording  of  the  Rule  B  will  be  changed  to  that  of 


=  tractive  power. 


RULE  3. 

Square  of  diameter  of  one  piston  x  mean  effective  steam  pressure  per  square  inch  x  stroke  in  feet 

Diameter  of  driving  wheel  in  feet 

Iii  ordinary  language,  Rule  3  would  read:  multiply  together  the  square  of  the 
diameter  in  inches  of  one  piston,  the  mean  effective  steam  pressure  per  square  inch 
and  the  length  in  feet  of  one  stroke.  The  product  thus  obtained,  divided  by  the 
diameter  in  feet  of  one  wheel,  will  be  the  tractive  power. 

EXAMPLE  6. — What  is  the  tractive  power  of  a  locomotive  whoso  cylinders  are  17 
inches  in  diameter  and  24  inches  stroke?  The  mean  effective  steam  pressure  is  90 
pounds  per  square  inch,  and  the  driving  wheels  60  inches  in  diameter. 


17  x  17  x  90  x  2 


=  10404  =  tractive  power. 


If  the  tractive  power  had  been  calculated  according  to  Rule  A,  the  result  would 
have  been  the  same.  But  Rule  3  is  evidently  the  simplest,  and  a  great  amount  of  time 
and  labor  will  be  saved  by  using  it. 

All  the  figures  expressing  the  adhesion  in  pounds,  as  shown  in  column  3  in  all  the 
following  tables,  have  been  found  according  to  Rule  3 : 


TABLE    5. 

EIGHT-WHEELED   LOCOMOTIVES. 


Cylinder*. 
Diameter.    Stroke. 

Diameter  of 
Driving  Wheels. 

AtllicM'.n. 

Weight  on  Drivers. 

Weight  on  Truck. 

Total  Weight. 

Hauling  Capacityon  Level  Track 
In  Tons  of  2,000  Pounds,  in- 
cluding Tender. 

Column  1. 

Column  2. 

Column  3. 

Column  4. 

Column  5. 

Column  <;. 

Column  7. 

Inch1  -. 

Inches. 

Lbe. 

Lbe. 

Lb*. 

Los. 

in    •    2H 

4.1  to  51 

4000 

90000 

101  II  III 

80000 

533 

U  x  22 

45  to  51 

5324 

26620 

18810 

39930 

709 

]•_'    •    22 

48  to  54 

5940 

2!>700 

1  1  *.-,(! 

44550 

792 

13  x   22 

49  to  57 

6828 

3414(1 

17(171) 

51210 

BIO 

14  x  24 

f>.r.  to  61 

Tti'.iT 

384K.-) 

L9242 

57727 

1026 

15  ,   21 

.V>  to  lili 

8836 

441  HO 

221  I'll) 

<>(i270 

1  17s 

16  x  24 

r.s  to  66 

BBSS 

4706;-) 

23832 

71497 

1271 

17  x  24 

60  to  66 

1(14114 

S30X 

M01Q 

78080 

1887 

18  x  24 

6i  to  'it; 

11472 

57380 

28680 

86040 

1529 

in  columns  1  and  2  an-  tin-  s:iine  ;is  those  in  Table  1. 
Figures  in  column  3  arc  ol>taiue<l  acrordiiix  to  Kule  3. 


18 


M<H>EI;\ 


Figures  in  column  4  are  obtained  by  multiplying  the  figures  in  column  :j  by  5. 
Figures  in  column  5  are  obtained  by  multiplying  the  figures  in  column  4  by  .5. 

Figures  in  column  6  are  obtained  by  adding  the  figures  in  column  4  to  those  in  column  5,  and  are  the  weight 
of  engines  in  working  order  with  water  and  fuel. 

Figures  in  column  7  are  obtained  by  dividing  the  figures  in  column  3  by  7£. 


TABLE    (>. 
MOGUL    ENGINES. 


Cylinders. 
Diameler.    Stroke. 

Diameter  of 
Driving  Wheels. 

Adhesion. 

Weight  on  Drivers. 

Weight  on  Trucks. 

Total  Weight. 

Hauling  Capacity  on  Level  Track 
in  Tom  of  2,COO  Pound*,  in- 
cluding Tender. 

Column  1. 

Column  2. 

Column  3. 

Column  -1. 

Column  5. 

Column  6. 

Column  1. 

Inches. 

Inches. 

Lbs. 

Lbs. 

Lbs. 

Lbs. 

11   x  16 

35  to  40 

4978.2 

24891 

4978 

29869 

663 

12  x  18 

36  to  41 

6480 

;ii'40o 

6480 

38880 

864 

13  x  18 

37  to  42 

7399.4 

36997 

7399 

44396 

986 

14  x  20 

39  to  43 

9046 

45230 

9046 

54276 

1206 

15  x  22 

42  to  47 

10607 

53035 

1(101)7 

63642 

1414 

16  x  24 

45  to  f>l 

12288 

61440 

1  22SS 

73728 

1638 

17  x  24 

49  to  54 

12739.5 

63697 

1  2739 

76436 

1698 

18  x  24 

51  to  56 

13722.3 

68611 

13722 

&i:u:\ 

1828 

19  x  24 

54  to  60 

14440 

72200 

14440 

86640 

1925 

Figures  in  columns  1  and  2  are  the  same  as  those  in  Table  2. 
Figures  in  column  3  are  obtained  according  to  Rule  3. 

Figures  in  column  4  are  obtained  by  multiplying  the  figures  in  column  3  by  5. 
Figures  in  column  5  are  obtained  by  multiplying  the  figures  in  column  4  by  .2. 

Figures  in  column  6  are  obtained  by  adding  the  figures  in  column  4  to  those  in  column  5,  and  are  the  weight 
of  engines  in  working  order  with  water  and  fuel. 

Figures  in  column  7  are  obtained  by  dividing  the  figures  in  column  3  by  7£. 


TABLE    7. 
TEN-WHEELED  ENGINES. 


Cylinders. 
Diameter.    Stroke. 

Diameter  of 
Driving  Wheels. 

Adhesion. 

Weight  on  Drivers. 

Weight  on  Truck. 

Total  Weight, 
with  Water 
and  Fuel. 

HanlingCapacitvon  Level  Track 
in  Tons  of  2,000  Pounds,  in- 
cluding Tender. 

Column  1. 

Column  2. 

Column  3. 

Column  4. 

Column  5. 

Column  6. 

Column  7. 

Inches. 

Inches. 

Lbs. 

Lbs. 

Lbs. 

Lbs. 

12  x  18 

39  to  43 

5981.5 

29907 

9570 

39477 

797 

13  x  18 

41  to  45 

6677.5 

33387 

10683 

44070 

890 

14  x  20 

43  to  47 

8204.6 

41023 

13127 

54150 

1093 

15  x  22 

45  to  50 

9900 

49500 

15840 

65340 

1320 

16  x  24 

48  to  54 

11520 

57600 

18432 

71:0:12 

1536 

17  x  24 

51  to  56 

12240 

61200 

1  9584 

80784 

1632 

18  x  24 

51  to  56 

13722.3 

68611 

21955 

90568 

1829 

19  x  24 

54  to  60 

14440 

721200 

23104 

95304 

1925 

Figures  in  columns  1  and  2  are  the  same  as  those  in  Table  3. 
Figures  in  column  3  are  obtained  according  to  Rule  3. 

Figures  in  column  4  are  obtained  by  multiplying  the  figures  in  column  3  by  5. 
Figures  in  column  5  are  obtained  by  multiplying  the  figures  in  column  4  by  .32. 

Figures  in  column  6  are  obtained  by  adding  the  figures  in  column  4  to  those  in  column  5,  and  are  the  weight 
of  engines  in  working  order  with  water  and  fuel. 

Figures  in  column  7  are  obtained  by  dividing  the  figures  in  column  3  by  7A. 


MI>IU:I;\  i.nm MI >n 1 i-:  t  <>\sn;i CTIOX. 


11) 


TABLE    8. 
CONSOLIDATED    ENGINES. 


rvlinilere. 
Diani.-u-r.    Stroke. 

Dianu-U'r  of 
Driving  WhoclB. 

Aillii-iiin. 

Weight  on  Drivers. 

Weight  on  Truck. 

Total  Weight, 
with  Waivr 
:ind  Fuel. 

Haulm);  Capacity  on  Level  Track 
in  Tons  of  iOIX)  Pounds,  in- 
cluding Tender. 

Column  1. 

Column  S. 

Column  3. 

Column  4. 

Column  5. 

Column  6. 

Column  7. 

Inches. 

Inches. 

Lbe. 

UN. 

Lb«. 

LOB. 

14  x  16 
i:>  x  18 
•Jn  x  24 
21'    -    I'l 

:;r.  to  38 

36  to  38 

4S  to  50 
.in  to  52 

7840 
10128 
18000 

20908.8 

88200 

5061;.-) 

!M)000 
104544 

6272 
8100 
14400 
16727 

45472 
58725 
104400 
121271 

1045 
1390 

24(10 
2787 

Figures  in  columns  1  and  2  are  the  same  as  those  in  Table  4. 
Figures  in  column  3  are  obtained  according  to  Rule  3. 

Figures  in  column  4  arc  obtained  by  multiplying  the  figures  in  column  3  by  5. 
Figures  iu  column  5  are  obtained  by  multiplying  the  figures  in  column  4  l»y  .16. 

Figures  in  column  6  are  obtained  by  adding  the  figures  in  column  4  to  those  in  column  5,  and  are  the  weight 
of  engines  in  working  order  with  water  and  fuel. 

Figures  in  column  7  are  obtained  by  dividing  the  figures  in  column  3  by  7|. 


CHAPTER    II. 

CONSTRUCTION  OF  CYLINDERS.— STEAM  PIPES.— SLIDE-VALVES. 

25.  The  general  practice  in  the  United  States  is  to  place  the  cylinders  outside  the 
frames  A,  A,  as  shown  in  Figs.  10  and  11 ;  but  on  further  examination  of  these  two 
figures,  we  find  that  there  is  a  considerable  difference  in  the  construction  of  the  cylin- 
ders, so  that  we  may  divide  these  into  two  classes.     In  one  class  we  may  place  the 
cylinder  with  half  the  saddle  cast  in  one  piece ;  and  in  the  other  class  we  may  place  the 
cylinder  with  the  saddle  cast  separate.     The  following  explanation  will  help  the  reader 
to  gain  a  clearer  understanding  of  the  difference  of  these  constructions. 

Fig.  10  shows  an  end  view  of  a  cylinder  with  half  the  saddle  cast  to  it.  In  this 
case  the  cylinder  casting  is  extended  over  the  frame  A  to  the  center  of  the  boiler,  and 
here  meets  a  similar  extension  from  the  opposite  cylinder  (not  shown  in  this  figure), 
the  two  being  bolted  together  by  the  bolts  I,  b,  b.  These  extensions  of  the  cylinders, 
or  those  parts  of  the  castings  which  extend  from  frame  to  frame,  constitute  the  cylin- 
der saddle ;  hence  this  type  of  cylinder  is  known  by  the  term  "  cylinder  and  half-saddle 
cast  in  one." 

26.  Fig.  11  shows  a  locomotive  cylinder  belonging  to  the  second  class,  in  which 
the  saddles  are  cast  separate,  the  cylinder  being  bolted  to  the  saddle  by  the  bolts  C,  C; 
the  manner  of  fastening  these  to  the  frames  is  similar  to  that  of  the  former  cylinders. 

Cylinders  with  half-saddle  cast  in  one  are  generally  used,  because  only  one  pattern 
is  needed  for  both  cylinders  in  a  locomotive ;  whereas,  the  saddle  being  cast  separately, 
we  need  a  greater  number  of  patterns. 

27.  In  this  chapter  we  will  consider  only  the  cylinders  with  half-saddle  cast  in  one 
because  they  are  the  most  popular  ones.     The  arrangement  of  the  steam-ways,  steam 
passages,  exhaust  passages,  as  well  as  all  other  details  of  these  cylinders,  are  shown  in 
Figs.  12,  13,  and  14. 

Fig.  12  shows  a  section  lengthways  of  the  cylinder,  the  section  being  taken  through 
the  line  a,  b,  drawn  in  Fig.  14. 

Fig.  13,  right-hand  side,  shows  a  section  of  the  cylinder  and  saddle  taken  through 
a  line  c,  rf,  drawn  in  Fig.  14. 

Fig.  13,  left-hand  side,  shows  an  outside  end  view  of  the  cylinder  and  half-saddle. 

Fig.  14  shows  a  plan  of  the  cylinders  and  saddle ;  the  right-hand  side  of  this  figure 
shows  a  section  of  half  the  saddle,  the  section  being  taken  through  the  line  c,  f,  drawn 
in  Fig.  13.  Similar  letters  in  these  three  views  indicate  the  same  pieces  or  parts  of  the 
cylinder  and  saddle. 

In  Fig.  12  the  cylinder  heads,  piston,  and  piston-rods  are  shown,  but  these  h.-ivc 
been  omitted  in  all  the  other  figures. 


\min-:i; v  i.<H-».Mnri\'E  ro\sTi;rrrin\. 


21 


.1,  .1  are  sections  of  the  frames  to  which  the  cylinders  are  firmly  bolted. 

/.'  represents  the  cylinder. 

('  represents  the  back  cylinder  head. 

1)  represents  the  front  cylinder  head. 

E  represents  the  piston. 

/•'represents  the  piston-rod. 

(i  represents  the  piston-rod  gland. 

//  represents  steam-chest  seat,  that  is,  the  surface  on  which  the  steam-chest  rests 


I  represents  the  valve  seat,  that  is,  the  surface  on  which  the  slide-valve  moves 
backward  and  forward. 

J represents  the  steam  passage.  This  steam  passage  terminates  at  01 nd  (that 

is,  the  end  in  the  saddle),  with  a  round  hole,  as  shown  at  J\.  Before  this  passage 
reaches  the  other  end,  it  is  divided  into  two  brandies,  each  one  terminating  in  the 


22  MODJSRN  LOCOMOTIVE   CONSTRUCTION. 

steam-chest  seat  with  the  openings  marked  ./2,  J.%  an(l  both  of  these  openings  lie  inside 
the  steam-chest  (the  steam-chest  is  not  shown  in  these  illustrations). 

28.  By  dividing  this  steam  passage  into  two  branches  three  advantages  are  gained. 
First,  the  steam  will  be  delivered  at  each  end  of  the  steam-chest,  so  that  the  steam  can 
freely  enter  the  steam  ports.     Second,  a  right  and  left-hand  cylinder  pattern  will  be 
avoided,  and  only  one  pattern  needed  for  both  cylinders.     Third,  if  the  cylinders  are 
accurately  planed  and  fitted  to  gauges,  they  will  be  interchangeable ;  that  is,  we  can 
use  such  a  cylinder  for  either  side  of  the  engine. 

29.  The  duty  of  the  steam  passage  is  to  conduct  the  steam  into  the  steam-chest ; 
the  steam  enters  the  opening  Jlt  and  is  delivered  into  the  steam-chest  through  the  open- 
ings J2,  J2.    This  is  the  only  duty  the  steam  passages  have  to  perform,  and  consequently 
the  steam  in  these  passages  will  always  flow  in  one  direction,  as  shown  by  arrows  2. 

30.  K  represents  the  exhaust  passage.     This  terminates  with  one  opening  in  the 
valve  seat,  and  this  opening,  marked  Klt  is  called  the  exhaust  port ;  the  other  end  ter- 
minates with  a  round  opening  K2,  a  little  above  the  saddle.     Some  designers  make  the 
form  of  this  opening  a  semicircle ;  others,  again,  will  make  it  a  square  or  oblong ;  and 
which  of  these  forms  is  to  be  adopted  will  depend  greatly  upon  the  judgment  and 
fancy  of  the  designer.     The  duty  of  the  exhaust  passage  is  to  conduct  the  steam  out  of 
the  cylinder  after  it  has  performed  its  work.     In  this  passage  the  steam  will  always 
flow  in  a  direction  as  indicated  by  arrow  3. 

31.  On  each  side  of  the  exhaust  passage,  Fig.  12,  is  a  channel  or  passage  marked 
L  L ;  these  passages  are  called  the  steam- ways.    For  the  sake  of  distinction  we  shall  call 
the  steam  way  nearest  the  front  cylinder  head,  the  "  front  steam- way,"  and  the  other 
one,  the  "  back  steam- way."    These  steam- ways  terminate  with  the  openings  L^  Ll  in 
the  valve  seat,  and  these  openings  are  called  the  steam  ports.    The  steam-ways  have  a 
double  duty  to  perform,  namely,  they  must  conduct  the  steam  into  the  cylinder,  and 
after  the  steam  has  performed  its  work,  they  must  conduct  the  steam  out  of  the 
cylinder.    For  instance,  when  the  piston  stands  in  the  position  as  shown  in  Fig.  12, 
the  steam  will  be  conducted  through  the  front  steam-way  into  the  cylinder  space 
between  the  front  cylinder  head  and  piston,  and  the  steam  will  flow  through  this 
steam- way  in  a  direction  as  shown  by  arrows  4.     On  the  other  side  of  the  piston  the 
steam  is  conducted  out  of  the  cylinder,  flowing  through  the  back  steam-way  in  a 
direction  as  indicated  by  arrows  5. 

But  now,  when  the  piston  stands  near  the  back  cylinder  head,  then  the  back  steam  - 
way  will  conduct  the  steam  into  the  cylinder,  the  steam  flowing  in  an  opposite 
direction,  to  that  shown  by  the  arrows  5,  and,  in  the  meantime,  the  front  steam-~\v;iv 
will  conduct  the  steam  out  of  the  cylinder,  causing  it  to  flow  in  an  opposite  direction 
to  that  indicated  by  the  arrows  4. 

Here,  then,  the  reader  will  perceive  what  was  meant  by  saying  the  steam-ways 
have  a  double  duty  to  perform.  We  draw  particular  attention  to  this  fact,  as  we  shall 
allude  to  it  again. 

The  metal  or  small  bars  marked  M,  between  the  steam  ports  and  the  exhaust  port, 
are  called  bridges. 

32.  N,  N  represent  the  cylinder  cocks.     0  represents  the  boiler,  or  to  be  more 
precise,  the  smoke-box  of  the  boiler,  to  which  the  cylinders  are  firmly  bolted  as  shown. 


ro.\sri:r<-ri<>\.  23 

PISTON   AND   ENGINE  CLEARANCE. 

33.  Piston  clearance  is  the  distance  /'  between  the  piston  and  cylinder  head  at  the 
end  of  a  stroke.     In  locomotive  cylinders  it  varies  from  £  to  £  inch;  it  is  generally 
equal  to  I  of  an  inch.     The  term  "  piston  clearance"  must  not  be  confounded  with  the 
term  "engine  clearance,"  or  simply  clearance,  which  is  the  space  between  the  piston 
and  head  plus  the  volume  of  the  steam-way,  or  we  may  say,  that  the  engine  clearance 
is  the  whole  minimum  space  between  the  piston  and  the  valve  face  at  the  completion 
or  beginning  of  a  stroke. 

COUNTERBORE. 

34.  The  cylinder  is  counter-bored  at  each  end,  generally  J  of  an  inch  larger  in 
diameter  than  the  bore  of  the  cylinder.     The  depth  of  the  couuterbore  should  be  such, 
that  when  the  piston  stands  at  the  end  of  a  stroke,  as  shown  in  Fig.  1'2,  the  packing 
ring,  //,  will  not  project  more  than  $  of  an  inch  beyond  the  edge  of  the  couuterbore,  or 
when   the  piston  stands  at  the  opposite  end  of  the  stroke,  the  packing  ring,  /(,  will  not 
project  more  than  i  of  an  inch  beyond  the  edge  of  the  corresponding  counterbore.     In 
any  case  care  must  be  taken  to  regulate  the  depth  of  the  couuterbore  so  as  not  to  allow 
the  whole  width  of  any  of  the  packing  rings  to  pass  over  it.     It  will  be  readily  under- 
stood that  should  any  of  the  packing  rings  travel  beyond  the  edge  of  the  couuterbore, 
they  will  at  once  adjust  themselves  to  the  larger  diameter,  and  thus  prevent  the  piston 
from  returning  without  doing  considerable  damage. 

:!.">.  For  the  cylinders  such  east-iron  should  bo  selected  as  will  wear  well  and 
equally;  it  must  be  hard  and  homogeneous,  yet  not  so  hard  as  to  prevent  the  tools 
from  cutting  during  the  process  of  boring,  planing,  drilling,  and  turning. 

.'!»;.  The  joints  beween  cylinder  and  cylinder  heads  are  made  metal  to  metal,  and 
ground;  the  part  to  be  ground  is  allowed  to  project  a  little  beyond  the  face  of  the 
flange,  as  shown  in  Figs.  12  and  14. 

The  bolts  securing  the  cylinder  heads  to  the  cylinder  are  usually  placed  from  4£ 
to  :>A  inches  from  center  to  center,  and  these  distances  will  determine  the  number  of 
bolts  to  be  used.  Their  diameter  should  be  such  that  the  stress  brought  upon  them  by 
the  steam  pressure  alone  will  not  exceed  5,000  pounds  per  square  inch. 

CYLINDER   PROPORTIONS   AND   DETAILS. 

37.  To  find  the  diameter  of  a  cylinder-head  bolt  we  must  first  know  the  initial 
steam  pressure  in  the  cylinder,  that  is,  the  steam  pressure  at  the  beginning  or  near  the 
beginning  of  the  stroke.  We  believe  that  12u  pounds  per  square  inch  for  the  initial 
steam  pressure  \\-jH  agree  very  closely  with  ordinary  practice,  but  the  tendency  is  to 
work  with  higher  steam  pressure  than  this.  Assuming  that  1'JO  pounds  per  square 
inch  for  the  initial  pressure  is  correct,  the  diameters  of  the  bolts  for  the  cylinder  head 
can  easily  be  found. 

For  example,  let  it  be  required  to  find  the  diameter  of  tl ylinder-head  bolts,  the 

cylinder  being  l<i  inches  in  diameter,  and  the  bolts  to  be  placed  about  ~>  inches  from 

center  to  center. 


24 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


The  area  of  a  circle  16  inches  in  diameter  is  201  square  inches;  multiplying  this 
area  by  the  initial  steam  pressure  per  square  inch,  we  have  201  x  120  =  24,120  pounds ; 
this  is  the  force  which  the  combined  strength  of  all  the  bolts  in  one  cylinder  head  must 
be  capable  of  resisting,  independent  of  the  stress  that 
is  already  placed  on  these  bolts  by  the  use  of  the  screw 
wrench.  Placing  these  bolts  about  5  inches  from  cen- 
ter to  center,  we  find  that  twelve  bolts  are  needed. 
Dividing  the  24,120  pounds  by  12,  we  have  ^^-  =  2,010 
pounds.  This  means  that  each  bolt  must  be  capable 
of  resisting  a  pulling  force  of  2,010  pounds. 

Tlio  area  of  the  cross-section  of  a  bolt  must  be  such 
that  each  square  inch  will  be  subjected  to  a  stress  of 
not  more  than  5,000  pounds.  But  now  the  pulling 
force  on  the  bolt  is  less  than  5,000  pounds,  hence  the 
area  will  also  be  proportionately  less  than  one  square 
inch.  Therefore,  to  find  the  area  of  cross-section  of 
the  bolt  we  divide  2,010  by  5,000,  and  ffcro  =  -4  5  this 
means  that  the  area  of  the  cross-section  of  the  bolt 
must  be  TO  of  a  square  inch,  and  consequently  will  be 
very  nearly  $  of  an  inch  diameter.  Adding  to  this 
diameter,  twice  the  depth  of  thread,  we  find  that  the 
bolts  for  a  16-inch  cylinder  head  should  be  &  of  an  inch 
in  diameter. 

38.  Some  master  mechanics  object  to  bolts  in  the 
cylinder  head,  and  demand  studs  in  place  of  them. 
Their  objection  to  bolts  is,  that  in  case  a  bolt  should 
break,   the  whole  cylinder  lagging  (marked  _R,  R  in 
Figs.  12  and  13)  must  be  taken  off  in  order  to  place  a 
new  bolt  in  position.     On  the  other  hand,  should  a 
stud  break,  that  part  left  in  the  cylinder  flange  can  be 
drilled  out,  without  disturbing  the  cylinder  lagging, 
and  hence  the  preference  for  studs. 

39.  The  cylinder  lagging  consists  of  strips  of  ordi- 
nary pine  fitted  around  the  cylinder,  the  thickness  of 
these  strips  filling  the  whole  space  between  the  body 
of  the  cylinder  and  the  outside  of  the  flanges.     The 
writer  believes  that  a  few  thicknesses  of  asbestos  paper 
placed   around   the  cylinder,  and  then  the  remaining 
space  filled  with  wood,  as  shown  in  Figs.  12  and  13,  is 
the  best  practice. 

40.  Figs.  15, 16, 17, 18,  and  19  show  the  thicknesses 
of  metal  in  locomotive  cylinder  heads,  cylinders,  and 

their  flanges.  These  dimensions  have  been  obtained  by  actual  measurements  of  the 
metal  in  cylinders  belonging  to  acknowledged  first-class  modern  locomotives,  and 
suitable  for  an  initial  pressure  of  120  pounds  per  square  inch.  In  these  figures  it 


Moi>Ki;\  /.(icoMorirK  coxsrari'Tiox.  25 

will  be  noticed  that  the  cylinder  flanges  are  considerably  thicker  than  those  of  the 
cylinder  head;  this  is  certainly  a  good  practice,  because  in  case  a  fracture  should  take 
place  through  some  obstruction  between  cylinder  head  and  piston,  the  break  will 
occur  in  the  cylinder  head,  and  not  in  the  cylinder  flange;  and  thus,  on  account 
of  the  comparative  cheapness  with  which  a  cylinder  head  can  be  replaced,  costly 
repairs  and  vexing  delays  wih1  be  avoided,  such  as  are  sure  to  follow  when  the  cylin- 
der flange  is  injured. 

41.  In  order  to  insure  a  good  cylinder  casting,  the  thickness  of  the  bridges  should 
be  about  the  same  as  that  of  the  cylinder  barrel,  and  the  thickness  of  metal  around  the 
exhaust  passages,  the  steam  passages,  and  sides  of  the  saddle  may  be  made  about  ^,  of 
an  Inch  less  for  the  smaller  cylinders,  and  about  £  of  an  inch  less  for  the  larger  cylin- 
ders. 

4±  It  sometimes  occurs  in  connection  with  railroads,  that  ferry-boats  must  be 
used  in  which  the  cylinders  are  very  large  in  diameter.  To  determine  the  thickness  of 
metal  in  these  cylinders,  the  following  facts  must  betaken  into  consideration :  first,  the 
thickness  of  metal  must  be  such,  that  when  the  cylinder  is  subjected  to  the  maximum 
steam  pressure,  fracture  cannot  take  place ;  second,  a  sufficient  amount  of  metal  must 
be  allowed  for  reboriug ;  third,  the  cylinder  must  be  sufficiently  stiff  to  prevent  jarring 
during  the  process  of  boring  and  planing ;  and  lastly,  cylinders  must  be  sufficiently 
strong  to  maintain  their  circular  form.  Rules  for  finding  the  thickness  of  metal  in 
cylinders  that  will  satisfy  all  the  foregoing  demands  have  been  given  by  a  number  of 
eminent  writers. 

The  following  rule  has  been  copied  from  J.  D.  VanBuren's  book  on  "  formulas  for 
the  strength  of  the  iron  parts  of  steam  machinery."  In  the  writer's  opinion  this  is  the 
best  rule,  and  will  always  give  the  proper  thickness  for  all  marine  or  stationary  steam- 
engine  cylinders.  For  locomotive  cylinders  the  thickness  found  according  to  this  rule 
will  be  rather  light  when  compared  with  the  thicknesses  given  in  Figs.  15,  16,  17,  18, 
and  19. 

EULE  4. — To  find  the  thickness  of  metal  in  cylinders :  Multiply  the  diameter  of 
the  cylinder  in  inches  by  the  steam  pressure  per  square  inch,  also  multiply  this  prod- 
uct by  the  constant  decimal  fraction  .0001;  add  to  this  last  product  the  square  root  of 
the  diameter  of  the  cylinder  in  inches  multiplied  by  the  constant  decimal  fraction  .15; 
the  result  will  be  the  thickness  of  metal  in  the  barrel  of  the  cylinder.  Or  we  may  write 
this  rule  thus : 


(Diam.  of  cyl.  in  inches  x  steam  pressure  persq.  inch  x  .0001)  +  .15  ^/diam.  of  cyl.  in  inches  =  thickness  of  cyl.  wall. 

EXAMPLE  7. — What  must  be  the  thickness  of  metal  in  the  barrel  of  cylinder,  49 
inches  in  diameter,  for  (ill  pounds'  steam  pressure  per  square  inch! 

(4!)  x  (JO  x  .0001)  +  .15  V49~=  1.34  inch  =  thickness  of  metal. 

Now  let  us  find,  according  to  rule,  the  thickness  of  metal  in  a  locomotive  cylinder 
16  inches  diameter.  Assuming  that  the  greatest  steam  pressure  these  cylinders  have 
to  resist  is  1'JO  pounds  per  square  inch,  we  have 

(16  x  120  x  .0001)  +  .15  v/Ui  =  .7920  of  an  inch, 


26 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


that  is,  the  thickness  of  metal  in  the  barrel  of  cylinder  should  be  fully  ;|  of  an  inch. 
Comparing  this  thickness  with  that  given  in  Fig.  16,  we  find  that  the  thickness  found 
according  to  this  rule  is  light;  in  locomotive  cylinders  it  should  have  been  oue.inch,  ;is 
shown.  By  a  little  reflection  we  can  discover  the  reason  why  the  thickness  of  metal  in 
locomotive  cylinders  should  be  more  than  in  cylinders  for  marine  or  stationary  engines. 
In  locomotive  cylinders  more  metal  must  be  allowed  for  reboring  than  in  cylinders  for 
the  other  classes  of  engines,  because  in  a  locomotive  the  piston  speed  is  generally  very 
high,  and  besides  this,  in  locomotive  engines  ashes  accumulate  around  the  exhaust 
nozzle  iu  the  smoke-box,  and  should  then  the  engine  be  in  motion  with  steam  shut  off 
(which  occasionally  occurs),  ashes  are  sometimes  drawn  in  the  cylinder  through  the 
exhaust  passage,  and  the  consequence  of  such  an  evil,  and  the  necessity  of  reboring,  can 
easily  be  conjectured.  Hence,  the  thickness  of  metal  in  a  cylinder  found  according  to 
this  rule  should  be  increased  from  £  to  J  of  an  inch  for  locomotives. 


STEAM    PORTS   AND   EXHAUST    POET. 

43.  Figs.  20  and  21  show  the  steam  ports  and  the  exhaust  port.  The  form  of  these, 
also  the  length,  breadth,  and  area  of  the  same,  are  now  to  be  considered  and  deter- 
mined. 

In  Fig.  21  it  will  be  noticed  that  the  ends  of  the  steam  ports  have  the  form  of  a 
semi-circumference  of  a  circle,  and  the  straight  lines  surrounding  the  exhaust  port 
are  joined  by  arcs,  whose  radii  are  equal  to 
those  of  the  semi-circumferences  of  the  ends 
of  the  steam  ports.  Ports  formed  in  this 
manner  are  superior  to  ports  with  square 
ends,  because  the  sliding  surface  of  the 
valve,  and  the  valve  seat  having  square- 


Fiff.20 

ended  ports,  are  liable  to  wear  into  grooves 
and  ridges,  particularly  at  the  angles.  Again, 
making  the  ends  of  the  ports  semicircular,  as  shown,  adds  strength  to  the  bridges. 
This  form  of  ports  will  also  enable  us  to  true  up  the  whole  port  with  a  milling  tool, 
and  facilitate  the  application  of  a  template  which  is  to  guide  the  milling  tool,  thus  pro- 
viding for  making  the  ports  in  all  cylinders  of  one  class  exactly  alike  in  regard  to  length, 
breadth,  and  distance  between  them ;  such  accuracy  is  a  matter  of  great  importance  in 
locomotive  engineering. 

The  length  of  a  steam  port  is,  within  certain  limits,  a  matter  of  choice.  The  length 
is  often  made  eqiial  to  the  diameter  of  the  cylinder,  sometimes  a  little  less,  but  should 
never  be  less  than  ?  of  the  diameter  of  the  cylinder.  When  the  length  of  a  steam  port 


LOCOMOTIVE   CONSTRUCTION.  27 

has  been  established,  then  its  breadth  must  be  such  that  the  port  will  have  the  proper 
area.  Here,  tlit'ii,  we  see  that  before  both  the  length  and  the  breadth  of  a  port  can  bo 
decided  upon,  tin-  area  of  the  port  must  he  known.  We  have  already  pointed  out  in  Art. 
31  that  the  steam-way  lias  a  double  duty  to  perform,  namely,  to  admit  steam  into  the 
cylinder,  and  to  conduct  it  out  of  the  same.  To  conduct  the  steam  out  of  the  cylinder 
requires  a  steam-way  of  greater  cross-section  than  that  which  simply  admits  the  steam; 
and  since  the  port  area  really  represents  the  cross-section  of  a  steam-way,  we  conclude 
that  the  port  area  for  conducting  the  steam  out  of  the  cylinder  must  be  greater  thau 
would  lie  required  for  admitting  the  steam.  The  reason  of  this  is  that,  when  the  steam 
is  admitted  into  the  cylinder,  the  pressure  of  the  steam  is  very  nearly  constant,  because 
the  pressure  is  sustained  by  the  How  of  steam  from  the  boiler,  and,  consequently,  the 
velocity  of  the  steam  will  be  nearly  constant.  But,  on  the  other  hand,  when  the  steam 
is  allowed  to  escape,  the  pressure  of  the  steam  is  generally  less  than  it  was  when  it 
entered  the  cylinder,  and  therefore  the  velocity  of  the  steam  will  be  slower.  Again,  as 
the  steam  continues  to  escape  the  pressure  in  the  cylinder  becomes  gradually  lower, 
and  consequently  the  velocity  also  decreases.  Now,  the  steam  must  be  discharged  as 
quickly,  or  nearly  so,  as  it  was  admitted;  but,  since  the  velocity  of  the  steam  is  slower 
when  it  (lows  into  the  air  than  when  it  flowed  into  the  cylinder,  the  area  of  the  steam 
port  for  the  release  of  steam  must  be  larger  than  the  area  that  would  be  required  for 
the  admission  of  steam.  Wo  therefore  conclude  that  if  the  area  of  a  steam  port  is  largo 
enough  for  the  release  of  steam,  it  will  always  bo  large  enough  for  the  admission  of 
steam. 

We  find  in  the  valuable  work  of  D.  K.  Clark  on  "Railway  Machinery,"  that  for  a 
piston  speed  of  000  feet  per  minute,  a  good  exhaust  will  be  obtained  when  the  area  of 
the  steam  port  is  ,'„  the  area  of  the  piston,  the  steam  being  in  an  ordinary  state  as  to 
dryness.  Assuming  that  for  a  slower  piston  speed  the  area  of  the  steam  port  must  be 
proportionately  less,  and  for  a  faster  piston  speed  proportionately  larger,  we  have  all 
the  data  necessary  to  find  the  area  of  the  steam  port  suitable  for  any  given  diameter  of 
cylinder  and  piston  speed.  Now,  since  for  a  piston  speed  of  GOO  feet  per  minute  the 
port  area  must  be  -fa  of  the  area  of  the  piston,  and  for  other  speeds  the  port  area  must 
be  in  proportion,  we  may  put  our  data  in  the  following  form: 

600  is  to  the  given  piston  speed  in  feet  as  -^  of  the  piston  area  in  inches  is  to  the  port 

area  ;    or  thus  : 
li(M)  :  the  given  piston  sp 1  in  feet  : :  ny  of  the  piston  area  in  inches  :  the  port  area. 

But  writing  our  data  in  this  form,  we  recognize  a  statement  of  the  simple  rule  of 
proportion ;  in  order,  then,  to  find  the  required  port  area,  we  must  follow  the  rule  of 
proportion,  consequently  we  have 

Rule  5. 

(liven  piston  speed  in  feet  per  minute  x  .1 

(.( M ,  =  port  area  in  fractional  parts  of  piston  area. 

In  ordinary  language  this  rule  would  read:  The  given  piston  speed  in  feet  per 
minute  multiplied  by  ,',,,  and  this  product  divided  by  GOO,  will  be  equal  to  the  port  area 
in  fractional  parts  of  the  piston  area. 


28 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


Then,  multiplying  the  area  in  square  inches  of  the  piston  by  the  port  area  found  by 
Eule  5,  we  obtain  the  number  of  square  inches  that  the  steam  port  area  must  contain. 

EXAMPLE  8.—  Find  the  steam  port  area  suitable  for  a  cylinder  17  inches  in 
diameter,  and  a  piston  speed  of  650  feet  per  minute. 

According  to  Rule  5,  we  have  — 

650^1 

600 
that  is,  the  port  area  must  be  equal  to 


. 

part  of  the  piston  area.     The  area  of  a 

piston  17  inches  in  diameter  is  226.98  square  inches,  hence  226.98  x  .108  =  24.51384. 
This  means  that  the  steam  port  area  must  be  equal  to  24£  square  inches  for  this  partic- 
ular piston  speed.  Of  course,  for  a  slower  piston  speed,  this  port  area  should  be  less. 
For  instance,  if  the  piston  speed  is  to  be  500  feet  per  minute,  and  the  diameter  of  the 
cylinder  17  inches,  as  before,  we  have  — 

500  x  .1 
~~600~~          '3; 

and  the  piston  area  226.98  x  .083  =  18.8  square  inches  for  port  area. 

With  the  aid  of  Rule  5  we  may  arrange  the  following  table,  showing  the  ratio  of 
port  area  to  piston  area.  This  table  will  be  found  usefiil  and  convenient,  because  with 
it  the  area  of  the  steam  port  can  be  found  with  less  time  and  labor. 


TABLE   9. 

PROPORTIONAL   STEAM   PORT  AREA. 


Speed  of  Piston  in  Feet  Per  Minute. 

Port  Area  in  Fractional  PartB  of  Piston  Area. 

200  

033 

250  

041 

300 

050 

350  

058 

400  

.    .      .066 

450 

075 

500  

083 

550  .  . 

.091 

600  

1 

650  

108 

700 

116 

RULE  6. — To  find,  with  the  aid  of  Table  9,  the  area  of  a  steam  port  suitable  for  a 
given  diameter  of  cylinder  and  a  given  piston  speed :  Multiply  the  area  in  square  inches 
of  the  piston  by  the  decimal  fraction  found  in  Table  9  on  the  same  line  that  the  given 
piston  speed  is  indicated,  the  product  will  be  the  steam  port  area  in  square  inches. 

EXAMPLE  9. — Find  the  area  of  a  steam  port  suitable  for  a  cylinder  14  inches  in 
diameter,  and  a  piston  speed  of  450  feet  per  minute. 

The  area  of  a  piston  14  inches  in  diameter  is  153.9  square  inches ;  referring  to  the 
piston  speed  of  450  feet  in  the  Table  9,  we  find  on  the  same  line  the  decimal  fraction 
.075,  hence  153.9  x  .075  =  11.5425 ;  that  is,  the  steam  port  area  should  be  equal  to  ll£ 
square  inches. 

We  have  already  stated  that  the  length  of  the  steam  port  should  be  equal,  or  very 
nearly  so,  to  the  diameter  of  the  cylinder,  and  not  less  than  £  of  the  same.  If,  then, 


'i-:  c()\xTi;r<-Ti<>.\ 


the  length  has  been  decided  upon,  we  simply  divide  the  area  of  the  steam  port  by  the 
length,  and  tin-  quotient  will  be  the  necessary  width  of  port.  Or,  if  we  are  compelled 
to  adopt  a  certain  width  for  the  steam  port,  we  divide  the  area  of  the  port  by  the 
width,  and  the  quotient  will  be  the  length  of  the  port.  In  these  calculations,  for  the 
sake  of  simplicity,  we  consider  the  ports  to  have  square  ends,  as  the  area  that  is  lost  in 
making  the  ends  circular  is  so  small  that  it  may  be  neglected. 

44.  The  length  of  the  exhaust  port  is  always  made  equal  to  that  of  the  steam  port. 
For  finding  the  breadth  of  the  exhaust  port,  a  graphic  process  has  been  here  adopted, 
and  such  as,  the  writer  believes,  will  be  easy  to  follow  and  understand.  Fig.  "22  shows 
a  section  of  a  slide-valve,  exhaust  and  steam  ports.  The  valve  is  shown  to  stand  in 
two  positions  on  the  valve  seat.  When  in  the  position  marked  A  the  valve  stands  in 
a  central  position,  that  is,  midway  of  its  extreme  travel,  and  when  in  position  U  it 


l< Lcnytlt-of-l-alve-seat 

Fig.22 

stands  at  one  end  of  its  extreme  travel.  In  this  latter  position  we  notice  that  the 
opening  of  the  exhaust  port  is  considerably  contracted,  and  should,  then,  the  opening 
C  be  much  smaller  than  the  width  of  the  steam  port,  the  free  escape  of  the  exhaust 
steam  would  be  interfered  with. 

We  consider  it  to  be  good  practice  to  make  the  exhaust  port  wide  enough  so  that 
when  the  valve  stands  in  an  extreme  position,  as  at  B,  the  opening  C  will  then  be  equal 
to  the  width  of  the  steam  port.  Therefore,  to  find  the  width  of  the  exhaust  port,  first 
indicate  on  paper  the  width  of  the  steam  port  and  the  thickness  of  the  bridge  (see  Fig. 
23).  On  these  draw  a  portion  of  the  slide  valve  in  an  extreme  position,  as  shown,  and 
then  make  the  exhaust  port  wide  enough  so  that  the  opening  C' will  be  equal  to  the 
width  of  the  steam  port.  Generally,  when  this  rule  is  followed,  the  width  of  the 
exhaust  port  will  be  equal  to  about  twice  that  of  the  steam  port  for  locomotive  engines ; 
hence  it  is  often  said,  that  the  width  of  the  exhaust  port  should  be  equal  to  twice  that 
of  the  steam  port  for  all  engines.  This  statement  should  be  received  with  caution,  as 
it  might  lead  to  error.  Following  the  graphic  method  here  explained,  satisfactory 
results  will  always  be  obtained. 

l~i.  The  valve  seat  for  locomotive  cylinders  is  generally  raised  one  inch  above  the 
surrounding  surface,  so  as  to  allow  for  wear. 


30 


MODEKX  LOCOMOTIVE   CONSTRUCTION. 


The  length  of  the  valve  seat  (see  Fig.  22)  should  be  such  that  the  valve  may  con- 
siderably overshoot  it  at  each  end  of  the  travel  when  in  full  gear ;  this  will  promote 
uniformity  in  wear,  but  care  must  be  taken  not  to  make  the  valve  seat  too  short, 
because  then  the  steam  would  pass  underneath  the  valve  into  the  steam  port. 

STEAM   PIPES. — STEAM   AND   EXHAUST   PASSAGES. 

46.  Figs.  24  and  25  show  the  steam  pipes  and  the  manner  of  connecting  these  to 
the  cylinders.  The  steam  pipe  has  only  one  duty  to  perform,  namely,  conducting  the 


Section  of  Steam-pipe 
ii'jli  lite  line  tj  /,- 


Fig  24 


OD 

o       o  o        o 


o        o         o        o 


Fig.25 


steam  to  the  cylinder,  and  therefore  its  cross-sectional  area  is  made  less  than  the  area 
of  a  steam  port  through  which  the  steam  is  both  admitted  and  exhausted. 

It  has  been  foimd  that  for  a  piston  speed  of  600  feet  per  minute  good  results  will 
be  obtained  when  the  cross-section  area  of  the  steam  pipe  is  equal  to  .08  (that  is,  y^) 
of  the  area  of  the  piston ;  for  slower  piston  speeds  proportionally  less,  and  for  higher 
piston  speeds  proportionally  greater.  Consequently,  to  find  the  cross-sectional  area  of 
a  steam  pipe,  we  again  apply  the  rule  of  proportion,  thus : 

600  :  given  piston  speed  :  :  0.08  :  steam-pipe  area  in  fractional  parts  of  piston  area. 

From  the  foregoing  remarks,  we  can  establish— 

RULE  7. — Multiply  the  given  piston  speed  in  feet  per  minute  by  the  decimal  .08, 
and  divide  the  product  by  600;  the  quotient  will  be  the  cross-sectional  area  of  the 
steam  pipe  in  fractional  parts  of  the  piston  area.  Or,  writing  this  rule  in  the  form  of 
a  formula,  we  have — 

Given  piston  ^dper  minute  x  .08  =  steam.pipe  area  in  fractionai  parts  of  piston  area. 

Then,  multiplying  this  proportional  area  by  the  piston  area  in  square  inches,  we 
obtain  the  number  of  square  inches  in  the  steam-pipe  area. 


Mni>i-:it\ 


r».v.s-/7,vry/r>.v. 


31 


K  \AMPLE  10. — Find  tin;  cross-sectional  steam-pipe  area  suitable  for  a  cylinder  17 
inches  in  diameter,  and  a  piston  speed  of  500  feet  per  minuto. 
According  to  Rule  7,  wo  have — 

500  x  .08 


(iOO 


=  .0666 


This  moans  that  the  steam  pipe  area  must  bo  equivalent  to  r^rfj,  of  the  piston 
area.  The  area  of  a  piston  17  inches  in  diameter  is  220.98  square  inches;  there- 
fore, 22ti.98  x  .Ou'b'G  =  15.11b'8G8.  That  is,  the  cross-sectional  steam-pipe  area  should  be 
154  square  inches. 

With  the  aid  of  Kule  7,  the  following  table  has  been  arranged,  showing  the  ratio 
between  the  steam-pipe  area  and  the  piston  area  for  different  speeds. 

I'sing  this  table  when  the  steam-pipe  area  is  to  be  found,  time  and  labor  will  be 
saved.  To  find  the  cross-sectional  area  of  the  steam  pipe  with  the  aid  of  this  table,  we 
have1 — 

b'ri.K  8. — Multiply  the  area  of  the  piston  in  square  inches  by  the  decimal  fraction 
found  in  Table  10,  on  the  same  line  that  the  given  piston  speed  is  indicated  ;  the  prod- 
uct will  be  the  number  of  square  inches  in  the  cross-sectional  area  of  the  steam  pipe. 


TABLE   10. 

PUOPOUTIONAL,   STEAM-PIPE   AllEA. 


Speed  of  Piwton  in  Feet  Per  Minnie. 

st<-:iin-jiijir  Area  in  Fractional  Parti"  c>f  Piston 
Area. 

200 

0"6 

250  

033 

300  

04 

350  

046 

400  

053 

450  

06 

500  

066 

550  

073 

600  '.  .  .  . 

08 

(l.'iii  

086 

700  

093 

EXAMPLE  11. — Find  the  steam-pipe  area  suitable  for  a  cylinder  16  inches  in  diam- 
eter, and  a  piston  speed  of  451)  feet  per  minute. 

Referring  to  Table  10,  we  lind  on  the  same  line  with  the  given  piston  speed  of  450 
feet  the  decimal  .0<i.  The  area  of  a  piston  l(i  inches  in  diameter  is  I'd]  square  inches; 
therefore,  i>()l  x  .06  =  12.06.  That  is,  the  steam-pipe  area  should  contain  12  square 
inches. 

47.  In  Fig.  24,  the  area  of  the  opening  •/  of  the  steam  passage  should  be  the  same 
as  that  of  the  steam  pipe,  because  the  steam  passage  is  a  continuation  of  the  steam 
pipe.  Consequently,  when  we  know  the  area  of  the  steam  pipe,  we  also  know  the  area 
of  the  opening./,  and  when  we  know  the  area  of  this  opening,  its  diameter  is  easily 
found  by  referring  to  a  table  of  areas  of  cin-Ies,  or  by  one  of  the  simple  rules  of  men- 
suration. From  these  remarks,  we  may  convetly  infer  that,  for  instance,  a  cylinder  17 


32 


MODERN  LOCO  MOTIVE   CONSTRUCTION. 


inches  in  diameter  and  a  piston  speed  of  GOO  feet  per  minute  will  require  a  larger  ami 
in  the  opening  J  than  the  same  cylinder  would  require  for  a  piston  speed  of  500  feet 
per  minute ;  and,  since  the  piston  speed  is  not  the  same  in  all  locomotives,  we  would 
naturally  expect  to  find  a  number  of  core  boxes  of  different  size  for  each  cylinder,  so 
that  the  size  of  a  steam  passage  could  be  changed  in  a  cylinder  pattern  to  suit  some 
particular  piston  speed.  To  carry  out  such  a  system  would  require  too  great  a  variety 
of  patterns ;  and  to  avoid  this,  master  mechanics  and  manufacturers  generally  group 
the  cylinders,  according  to  their  diameter,  into  different  classes,  and  adopt  for  each 
class  some  particular  diameter  for  the  opening  J. 

In  the  following  table,  in  column  2  are  given  the  diameters  of  the  openings  J  for 
the  different  classes  of  cylinders,  such  as  are  generally  adopted ;  of  course,  some  makers 
will  vary  slightly  from  these  figures.  In  column  3  the  diameters  of  the  exhaust  open- 
ing (marked  K,  Fig.  24)  are  given.  If  these  exhaust  openings  are  made  square,  or  of 
some  other  form,  their  area  should  contain  about  the  same  number  of  square  inches  as 
the  circular  openings  corresponding  to  the  diameters  given  in  column  3,  Table  11. 


TABLE    11. 

SIZE   OF    STEAM    AND   EXHAUST   OPENINGS. 


Column  I. 

Column  2. 

Column  3. 

Diameter  of  Cylinders. 

Diameter  of  Steam  Opening  «/, 
in  Inches. 

Diameter  of  Exhaust  Opening  K. 
in  Inches. 

10 

3 

34 

11 

3 

3i 

12 

3i 

4 

13 

3i 

4 

14 

4i 

5 

15 

4i 

5 

16 

4± 

5 

17 

4} 

5 

18 

4i 

5 

19 

4f 

5 

20 

5 

5 

22 

5 

5 

i 

When  an  engine  is  to  be  designed  for  a  very  fast  speed,  we  would  advise  to  deter- 
mine the  area  of  the  opening  J  according  to  Rule  7,  and  not  follow  the  diameters  given 
in  the  last  table.  The  area  of  any  cross  section  of  the  steam  passage  should  contain 
the  same  number  of  square  inches  as  the  opening  J. 

In  regard  to  the  exhaust  passage,  good  results  will  be  obtained  when  the  area  of 
any  cross  section  in  the  neighborhood  of  the  line  a  I,  Fig.  24,  is  made  larger  than  the 
exhaust  opening  K;  in  fact,  we  have  always  obtained  good  results  by  making  it  as  large 
as  possible.  With  large  exhaust  passages  the  flow  of  exhaust  steam  will  not  be  so  irregu- 
lar as  when  smaller  passages  are  used.  In  the  writer's  opinion,  large  exhaust  passages 
will  improve  the  draft  of  an  engine,  and  to  some  extent  lessen  the  back  pressure  in 
the  cylinder.  No  rules  have  been  established  to  determine  the  area  of  the  exhaust 


33 

opening  K-    The  (liaint'ter  for  these  openings,  given  in  Table  11,  have  been  obtained 
by  actual  measurements. 

4S.  Often  it  will  lie  found  that,  in  designing  a  locomotive  cylinder,  the  space 
allotted  for  the  steam  passages  and  exhaust  passage  in  the  neighborhood  of  the  line  c  d, 
iMg.  -4,  is  very  small;  therefore,  great  care  and  good  judgment  must  I  e  used  to  ob- 
tain the  proper  cross-sectional  area  in  either  passage.  The  result  of  carelessness  right 
here  will  be  that  either  one  or  the  other  passage  is  too  small,  and  the  engine  will  fail 
to  do  the  work  that  it  was  intended  it  should  do. 

STEAM    PIPES. 

49.  Steam  pipes,  Figs.  '24  and  L!.">,  for  locomotives  are  made  of  cast-iron;  their 
thickness  of  metal  for  smaller  engines  is  about  £  of  an  inch,  and  for  larger  engines 
about  |  of  an  inch. 

On  account  of  some  practical  difficulties  that  must  be  overcome,  ordinary  flat 
joints  cannot  lie  used  between  the  T  pipe  and  the  steam  pipe,  neither  between  the  steam 
pipe  and  cylinder  saddle. 

The  first  difficulty  that  presents  itself  is  the  expansion  and  contraction  due  to  the 
great  change  of  temperature  to  which  the  steam  pipes  in  locomotives  are  exposed,  and 
therefore  we  must  adopt  a  joint  which  possesses  a  small  amount  of  flexibility. 

The  second  difficulty  that  presents  itself  is  of  a  practical  nature,  namely,  the 
impossibility  to  construct  a  boiler  and  fit  the  cylinder  saddle  to  the  outside  of  the 
smoke-box  with  absolute  accuracy,  yet  a  steam  pipe  of  proper  length  is  expected  to  fit  at 
once  iii  its  place,  without  any  more  labor  than  would  be  required  if  everything  elm  had 
Keen  perfectly  accurate;  hence,  the  joint  must  possess  a  small  amount  of  adjustability. 

Adopting  a  ball  joint,  the  foregoing  difficulties  can  be  overcome.  These  ball  joints 
are  made  (as  shown  in  Fig.  124)  by  interposing  a  brass  ring  between  the  T  pipe  and 
steam  pipe,  and  another  one  between  the  steam  pipe  and  cylinder  saddle.  Each  brass 
ring  has  a  spherical  and  a  flat  surface.  Now,  it  must  be  readily  perceived  that  with 
such  rings  interposed  the  steam  pipe  can  be  slightly  moved  up  or  down  or  sideways, 
and  still  maintain  a  steam-tight  joint.  This  kind  of  ball  joint  will  also  be  sufficiently 
flexible  to  allow  for  the  contraction  and  expansion  of  the  steam  pipe. 

PISTON    SPEED. 

51).  To  determine  the  piston  speed  in  feet  per  minute  according  to  the  following 
rule,  we  must  know  the  speed  of  train  in  miles  per  hour,  the  diameter  of  the  driving 
wheels  in  feet,  and  the  length  of  stroke  in  feet: 

KYuE  1).  —  To  find  the  piston  speed  in  feet  per  minute  in  a  locomotive,  multiply  the 
number  of  feet  in  a  mile  by  the  speed  of  train  in  miles  per  hour,  divide  the  product 
by  the  circumference  in  feet  of  the  driving  wheel  multiplied  by  (!0,  and  multiply  the 
quotient  by  twice  the  length  of  the  stroke  in  feet  ;  the  product  will  be  the  piston  speed 
in  feet  per  minute.  Or,  writing  this  rule  in  the  shape  of  a  formula,  we  have  — 


iXunibiT  of  fret  >        \  Spi'i'il  nf  train  in  I 
in  ;i  mil.-         \    '    I     mill's  JKT  lionr     \  i         piston   SIII-IM!         ) 

S  nr,.,,,,,f,.r,.,,  ......  f  .Irivi,,,?  wl.,,1  ,  *   twK>|-  "'"  S"'"k"  '"  1""t  '     /  in   f,,.   ,,,,-  min,,.,.  * 


'• 


34  MODERN  LOCOMOTIVE   CONSTRUCTION. 

EXAMPLE  12. — Find  the  piston  speed  in  feet  per  minute  in  a  locomotive  whose 
driving  wheels  are  5  feet  in  diameter ;  stroke,  2  feet ;  and  speed  of  train,  35  miles  per 
hour. 

According  to  Rule  9,  we  have — 

5280  x  35         _7o,fio 
15.7  x  60  X 

That  is,  the  piston  speed  will  be  784i60%  feet  per  minute.  In  order  to  assist  the  reader 
to  understand  the  foregoing  rule,  the  following  explanation  is  offered : 

First,  we  multiply  the  number  of  feet  in  a  mile  by  the  speed  of  train  in  miles  per 
hour ;  this  product  will  give  the  number  of  feet  the  locomotive  travels  in  one  hour, 
and,  since  there  are  5,280  feet  in  a  mile,  and  the  speed  of  train  in  our  example  is  35 
miles  per  hour,  we  have  5,280  x  35  =  184,800  feet  that  the  locomotive  travels  during 
one  hour.  To  find  the  number  of  feet  that  the  locomotive  travels  during  one  minute, 
we  divide  the  number  of  feet  per  hour  by  GO,  because  there  are  60  minutes  in  one  hour ; 
hence,  in  our  example,  -^r^1  =  3,080  feet ;  that  is,  during  one  minute  the  locomotive 
travels  through  a  distance  of  3,080  feet.  To  find  the  number  of  revolutions  of  the 
wheel  per  minute  (which  is  necessary  in  this  case),  we  divide  the  distance  traveled  per 
minute  by  the  circumference  of  the  driving  wheel.  In  our  example,  the  diameter  of 
the  driving  wheel  is  5  feet ;  the  circumference  of  such  a  wheel  is  15.7  feet ;  therefore, 

— —  =  196.17  number  of  revolutions  per  minute.  During  every  revolution  of  the  wheel 
.Lo.  t 

the  piston  travels  through  twice  the  length  of  the  stroke ;  therefore,  mutiplyiug  the 
number  of  revolutions  per  minute  by  twice  the  length  of  the  stroke,  the  piston  speed 
per  minute  will  be  obtained.  In  our  example,  the  stroke  is  2  feet;  therefore,  196.17  x 
4  =  784  j^o  feet-  That  is,  the  piston  speed  is  784-jVo  feet  per  minute.  » 

SLIDE-VALVES,  AND   MOVEMENT   OF   SLIDE-VALVES. 

51.  Slide-valves  are  sometimes  made  of  brass,  but  generally  of  cast-iron.  Cast- 
iron  slide-valves  are  more  durable  than  brass  valves,  but  the  latter  do  not  wear  the 
valve's  seat  as  quickly  as  the  cast-iron  valves. 

The  ordinary  form  of  slide-valve,  such  as  is  generally  used  in  locomotives,  is 
shown  in  Figs.  26  and  27.  Fig.  26  represents  a  cross-section  of  the  valve ;  one-half  of 


Fig.  26  Fi(J,  27  Fig.  28 

Fig.  27  shows  a  section  lengthwise  of  the  valve,  and  the  other  half  an  outside  view  of 
the  same.  The  thickness  of  metal  at  a  is  generally  made  1  in.,  and  at  b  about  i  in. 
The  sides  c  d,  e/are  extended  upwards  until  they  become  flush  with  the  top,  />,  of  the 
valve;  in  some  cases  these  sides  are  extended  a  little  beyond  the  top  of  the  valve. 
This  has  been  done  for  the  following  practical  reasons:  In  the  first  place,  a  large 


Moni':i;\  LOCOMOTIVE  cn\firiii'cTro2f. 


35 


surface  is  obtained  against  which  the  valve  yoke  can  bear.  Secondly,  this  form  of 
valve  can  be  laid  on  its  back,  and  thus  speedily  and  conveniently  secured  to  the  planer, 
when  the  valve  face  is  to  be  planed;  this  is  a  matter  of  no  small  importance  in  a  large 
locomotive  establishment  where  a  number  of  valves  have  to  be  planed  daily.  The 
recesses  g  g,  shown  in  Fig.  26,  are  simply  for  the  purpose  of  making  the  valve  as  light 
as  possible.  Some  master  mechanics  object  to  these  recesses,  because  they  believe  that 
they  will  hold  the  oil  (which  is  usually  admitted  through  the  top  of  the  steam  chest), 
and  prevent  the  oil  from  falling  upon  the  valve  seat,  and  thus  not  find  its  way  into  the 
cylinder.  For  this  reason  a  valve  has  been  adopted  having  a  form  as  shown  in  Fig. 
28.  This  form  of  valve,  although  used  on  some  roads,  has  not  been  favorably  received 
on  other  roads,  because  it  takes  up  too  much  room  in  height,  and  besides  it  is  an  incon- 
venient casting  to  fasten  to  the  planer  when  the  face  is  to  be  planed  or  replaned.  The 
writer  would  recommend  the  adoption  of  a  valve  having  a  form  as  shown  in  Fig.  26, 
and  believes  that  the  fear  of  the  recesses  g  g  preventing  the  oil  from  flowing  into  the 
cylinder  is  groundless,  and  that  the  constant  flow  of  steam  into  the  chest  will  not  allow 
the  oil  to  lay  still  on  any  part  of  the  valve. 

T)!'.  The  duty  of  the  slide-valve  is  to  control  the  flow  of  steam  into  and  out  of  the 
cylinder,  that  is,  the  valve  (as  its  name  implies)  slides  backward  and  forward  on  the 


Yoke  Brace 

Fig.29 
Valve  Gear 

for  an 
Eight  wheeled,  locomotive 

valvo  seat,  thus  opening  and  closing  the  steam  ports  at  proper  times.  Whether  it  will 
perform  this  duty  or  not,  depends  upon  the  form  and  motion  of  the  valve. 

Fig.  29  shows  a  complete  locomotive  valve  gear;  the  names  of  the  different  pieces 
of  the  mechanism  are  plainly  marked  on  the  drawing,  so  that  here  any  further  defini- 
tions of  these  pieces  will  be  unnecessary. 

53.  To  construct  a  slide-valve  and  assign  to  it  the  proper  motion,  such  as  shown 
in  Fig.  29,  may  seem  to  be  a  difficult  subject  for  solution;  and  so  it  would  be,  if,  ri.dit 
in  the  beginning,  we  do  not — wheresoever  we  can — throw  out  of  consideration  all  such 
pieces  of  mechanism  as  have  a  complicating  influence  upon  the  motion  of  the  valve. 
Hence  it  is  of  the  utmost  importance  first  to  reduce  this  subject  to  its  simplest  form. 
It  will  lie  noticed  that  tin-  operation  of  the  valve  is  controlled  by  two  eccentrics :  one 
eccentric  is  used  for  the  forward  motion  and  the  other  for  the  backward  motion  of  tin- 
engine.  Here  we  may  simplify  our  subject  by  leaving  out  of  consideration  the  back- 


36 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


ward  eccentric,  because  when  the  valve  is  made  to  accomplish  the  desired  results  with 
one  eccentric  its  form  will  not  have  to  be  changed  when  the  other  eccentric  is  added. 
But  leaving  the  backward  eccentric  out  of  the  question,  we  may  also  leave  the  link  out 
of  consideration,  because  the  link  only  serves  to  connect  the  two  eccentric-rods  so  as 
to  enable  the  engineer  to  put  wholly  or  partly  into  gear  one  or  the  other  eccentric. 
The  lifting-shaft  is  simply  used  for  moving  the  link  up  or  down  as  the  case  may  be ; 
and  since  the  link  has  been  thrown  out  of  consideration,  we  may  treat  the  lifting-shaft 
likewise.  The  rocker  is  simply  used  for  the  purpose  of  connecting  the  eccentric-rod  to 
the  valve-rod,  and  although  it  affects  the  position  of  the  eccentrics,  and  in  some  cases 
the  travel  of  the  slide-valve,  it  will  not  affect  the  laws  relating  to  the  construction  of 
the  valve,  and  therefore  we  also  throw  this  out  of  consideration. 

54.  Reducing  our  subject  as  described,  and  connecting  the  eccentric-rod  directly  to 
valve  stem,  we  obtain  a  simple  arrangement,  Fig.  30,  such  as  is  often  used  in  stationary 


B 


engines;  of  course,  in  this  arrangement  we  must  assume  that  the  driving  axle  of  the 
locomotive  is  represented  by  the  crank  shaft  C,  and  the  eccentric  placed  on  the  end  of 
the  shaft  as  shown.  In  this  arrangement,  simple  as  it  is,  a  feature  exists  which  has  a 
somewhat  complicating  influence  upon  the  motion  of  the  valve,  and  therefore  will 
interfere  with  the  simplicity  of  our  study  of  the  laws  relating  to  the  form  of  the  valve. 
The  feature  alluded  to  is  the  angle  that  the  eccentric-rod  makes  with  the  center  line,  A 
B.  This  angle  varies  during  the  travel  of  the  valve,  and  consequently  the  motion  of 
the  valve  will  be  slower  during  one  half  of  the  travel  than  during  the  other  half. 
Thus,  for  instance :  Let  the  line  A  B  in  Fig.  31  represent  the  line  A  B  shown  in  Fig. 
30.  The  circle  sl  tt  ult  Fig.  31,  will  represent,  in  an  exaggerated  manner,  the  path  of 


Fiy.31 


Fig.32 


the  center  x  of  the  eccentric,  shown  in  Fig.  30,  and  lastly,  the  distance  from  the  center 
x  to  the  center  t  of  the  eccentric-rod  pin  in  Fig.  30  is  represented  by  the  line  tl  t  in 
Fig.  31.  Now,  referring  only  to  Fig.  31,  when  the  valve  stands  in  an  extreme  position 
of  its  travel,  the  center  of  the  eccentric-rod  pin  will  be  at  »,  the  center  of  the  eccentric 
will  be  at  MI?  and  the  center  line  of  the  eccentric-rod  will  lie  in  the  line  A  B.  Again, 
when  the  valve  stands  in  the  other  extreme  position  of  its  travel,  the  center  of  the 
eccentric-rod  pin  will  be  at  s,  the  center  of  the  eccentric  will  bo  at  *„  and  the  center 
line  of  the  eccentric-rod  will  lie  in  the  line  A  B.  When  the  slide-valve  stands  central 


37 

that  is,  midway  between  the  extreme  ends  of  its  travel,  the  center  of  tiie  eccentric-rod 

j)in  will  be  nl  /,  exactly  midway  l»<'t\vt'cii  the  points  \  and  it. 

From  the  point  /  as  a  center,  and  with  a  radius  equal  to  the  distance  C  t,  describe 
an  an- ;  this  an-  will  intersect  the  circumference  *,  /,  «,  in  the  points  /,  and  t.,.  Join  the 
points  /  and  /,  l>y  a  straight  line,  also  draw  a  straight  line  from  the  point  t  to  the  point 
(.,;  then  the  straight  line  /  /,  or  /  /.,  will  represent  the  center  of  the  eccentric-rod  when 
the  valve  stands  in  a  midway  position  of  its  travel;  the  point  t  will  1  >e  the  center  of  the 
eccentric-rod  pin  and  the  point  /,,  or  the  point  (._,  will  he  the  center  of  the  eccentric. 
Assume  that  the  shaft  is  turning  in  the  direction  indicated  by  the  arrow.  When  the 
eccentric-rod  pin  has  traveled  from  n  to  /,  equal  to  half  the  travel,  the  slide-valve  has 
also  completed  one-half  of  its  travel,  and  the  center  of  the  eccentric  has  traveled 
through  the  arc  nl  /,.  Again,  during  the  time  that  the  eccentric-rod  pin  travels  from 
t  to  N,  equal  to  half  the  travel,  the  center  of  the  eccentric  will  travel  through  the  arc  tl 
Sj.  lint  now  notice  the  difference  of  length  of  the  two  arcs  tl  s}  and  t}  >it ;  this  plainly 
shows  that  the  eccentric-rod  pin  will  travel  slower  from  M  to  /  than  from  t  to  s,  and 
consequently  the  travel  of  the  valve  will  be  affected  likewise.  Or,  we  may  say,  that 
the  angle  formed  by  the  lines  /  /,,  and  ./  //  destroys  the  symmetry  of  the  valve  motion. 
Now,  in  the  study  of  the  laws  relating  to  the  motion  of  the  valve  and  the  duties  it  has 
to  pel-form,  such  a  motion  will  complicate  matters,  and  will  prevent  us  from  tracing 
the  action  of  the  valve  so  readily  as  when  both  halves  of  the  travel  are  described  in 
equal  times,  and  therefore  the  reader  will  perceive  the  necessity  of  changing  the  valve 
gear  to  one  which  will  give  the  valve  a  perfectly  symmetrical  motion. 

In  the  first  place,  it  will  be  easily  seen  that  the  longer  we  make  the  eccentric-rod 
—leaving  the  travel  of  the  valve  the  same — the  smaller  will  be  the  angle  between  the 
line  /  /,  (which  represents  the  center  of  eccentric-rod),  and  the  line  A  B,  and  conse- 
quently the  times  in  which  the  halves  of  the  travel  of  the  valve  are  described  will  be 
nearer  equal,  and  when  we  assume  the  eccentric-rod  to  be  of  an  infinite  length  the 
angle  will  vanish  and  each  half  of  the  travel  of  the  valve  will  be  described  in  equal 
times,  and  the  motion  will  be  symmetrical;  in  fact,  the  valve  will  have  precisely  the 
same  motion  as  that  obtained  with  a  valve  gear,  as  shown  in  Fig.  32,  to  which  we  shall 
now  call  attention. 

'>'>.  In  this  tigure,  in  place  of  using  an  eccentric-rod,  the  valve-stem  is  lengthened, 
and  to  its  end  a  slotted  cross-head  is  forged.  The  eccentric  has  also  been  dispensed 
with,  and  in  its  place  a  pin  //  fastened  into  the  end  of  the  crank-shaft  has  been  adopted. 
The  distance  between  the  center  ('  of  the  crank-shaft  and  the  center  of  the  pin  //  must 

always  be  equal  to  the  distance  between  th nter  ('and  the  center  x  of  the  eccentric 

shown  in  Fig.  oil.  This  distance  from  ('  to  ./  is  called  the  eccentricity  of  the  eccentric, 
and  is  equal  to  one-half  of  the  throw,  or  in  other  words  the  throw  of  an  eccentric  is 
equal  to  twice  its  eccentricity.  In  this  particular  case,  as  shown  in  Fig.  :>0,  the  throw 
is  equal  to  the  travel  of  the  valve;  by  the  travel  of  the  valve  is  meant  the  distance 
between  the  extreme  points  of  its  motion.  In  all  direct  acting  valve  motions,  that  is 
when  no  rocker  or  link  is  used,  the  throw  of  the  eccentric  will  be  equal  to  the  travel  of 
the  valve.  In  locomotives,  the  travel  of  the  valve  is  not  always  equal  to  the  throw  of 

th centric,  the  difference  being  due  to  the  inlluence  of  the  link,  and  often  to  the 

unequal  length  of  the  rocker-arms. 


38 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


ralty  Seat 


/Path  of  the  centre 
of  the  eccentric  (nee  Fig  30) 


Now,  turning  our  attention  to  Fig.  32,  we  notice  that  by  substituting  for  the 
eccentric  a  pin  y  in  the  end  of  the  crank-shaft,  we  really  adopt  a  crank,  and  this  we 
can  do  without  affecting  the  correctness  of  the  reasoning  relating  to  the  movement  of 
the  valve,  because  the  action  of  the  eccentric  is  precisely  the  same  as  that  of  a  crank 
whose  length  is  equal  to  half  the  throw ;  the  only  reason  why  eccentrics  are  adopted 
is  that  they  are  more  convenient  to  use ;  in  fact,  cranks  in  many  cases  cannot  be  used, 
the  peculiar  construction  of  the  machine  preventing  their  adoption ;  in  no  case  is  an 
eccentric  adopted  because  a  different  motion  to  that  due  to  a  crank  is  desired. 

We  have  drawn  particular  attention  to  this  fact,  because  some  mechanics  (a  good 
many  of  them)  have  a  misty  notion  of  the  action  of  an  eccentric. 

As  the  shaft  revolves  (see  Fig.  32)  the  pin  in  the  end  of  the  shaft  will  move  in  the 
slot  of  valve-stem's  cross-head,  and  thus  always  allowing  the  center  line  of  the  valve- 
stem  to  coincide  with  the  line  A  B. 
It  must  also  be  plain  that  as  the 
shaft  revolves  the  center  of  the  pin 
will  describe  a  circle,  and  it  is  the 
circumference  of  this  circle  that  will 
enter  into  the  solutions  of  the  fol- 
lowing problems.  Once  more,  the 
reader  will  readily  perceive  that  the 
length  of  the  valve-stem  will  in  no 

Dtameter  of  thU  circle- travel  of  ttemlve.  WiS6  affect    the    motion    Of    the  ValVB, 

hence  we  may  leave  this  also  out  of 

consideration,  and  place  the  circumference  of  the  circle  which  represents  the  path 
of  the  pin  y  on  the  end  of  the  slide-valve,  as  shown  in  Fig.  33.  Here,  then,  we  notice 
that  our  original  subject,  that  of  finding  the  proper  motion  and  form  of  a  valve,  a 
subject  in  which  all  the  different  pieces  of  mechanism  as  shown  in  Fig.  29  had 
entered,  has  been  reduced  and  simplified  to  that  having  only  the  pieces  of  mechanism 
as  shown  in  Fig.  33. 

THKEE   CONDITIONS   A   SLIDE-VALVE   MUST   FULFILL. 

56.  The  entrance  of  steam  into  the  cylinder  is  regulated  by  the  two  outside  edges, 
a  ft  and  c  d,  of  the  slide-valve,  Fig.  34 ;  the  exit  of  the  steam  is  regulated  by  the  two 
inner  edges,  efandg  h,  of  the  slide-valve;  and  the  correct  admission  and  exhaust  of 
the  steam  depends  upon  these  edges,  the  motion  of  the  valve — that  is,  the  travel  of 
the  valve — and  the  position  of  the  eccentric. 

All  slide-valves  must  be  capable  of  fulfilling  the  three  following  conditions,  and  if 
a  slide-valve  cannot  do  this,  the  engine  will  not  work  satisfactorily : 

First  Condition. — Steam  must  be  admitted  into  the  cylinder  at  one  of  its  ends  only 
at  one  time.  To  satisfy  this  condition,  the  length  of  the  valve  from  a  to  c  must  at  least 
cover  both  steam  ports,  when  the  valve  stands  in  a  central  position,  as  shown  in  Fig. 
34.  This  length  of  the  valve  cannot  be  less,  because  if  it  is  made  less,  steam  will  enter 
both  ends  of  the  cylinder  at  one  time,  and  consequently  bad  results  will  follow. 

Second  Coi////tii»i. — The  valve  must  allow  the  steam  to  escape  from  one  end  of  the 
cylinder,  at  least  as  soon  as  it  is  admitted  into  the  other  end  of  the  cylinder.  To  fulfill 


I.O<-OM<>TI\  /•: 


39 


this  second  condition,  the  length  of  the  exhaust  cavity  in  the  valve,  or,  in  other  words, 
the  distance  between  the  two  inner  edges,  efanAg  //,  Fig.  34,  must  be  equal  to  the  sum 
of  t lie  widths  of  the  two  bridges  added  to  the  width  of  the  exhaust  port.  The  length 
of  the  exhaust  cavity  in  a  valve  whose  outside  edges  just  cover  the  steam  ports,  must 
not  lie  made  less  than  shown  in  Fig.  34,  because  if  it  is  made  less,  steam  will  be  ad- 
mitted into  one  end  of  the  cylinder,  some  time  before  the  steam  in  the  other  end  of  the 
cylinder  is  released,  and  consequently  a  considerable  amount  of  back  pressure  will  be 
the  result.  When  the  outer  edges  of  a  valve  overlap  the  steam  ports,  such  as  shown 
in  Fig.  :!<>,  then  its  exhaust  cavity  can  be  made  less,  within  certain  limits,  and  still 
satisfy  the  second  condition. 

Tlni-il  Cnndititiii. — The  valve  must  cover  the  steam  ports  so  as  not  to  allow  the 
steam  to  escape  fi'oin  the  steam  chest  into  the  exhaust  port.     To  fulfill  the  third  con- 


Fi y.  3 


Fid-  35 


fig.  36 


dition,  the  length  of  the  exhaust  cavity,  that  is,  the  distance  between  the  edges  e  and  g, 
Fig.  34,  must  not  be  made  greater  than  the  sum  of  the  width  of  the  two  bridges 
added  to  the  width  of  the  exhaust  port,  in  a  valve  whose  outside  edges  just  cover  the 
steam  ports.  If  the  length  of  the  exhaust  cavity  in  such  a  valve  is  made  greater,  then 
the  distance  between  the  edges  a  and  e,  or  the  distance  between  the  edges  g  and  c  will 
be  less  than  the  width  of  steam  ports,  and  consequently  the  steam  will  be  permitted  to 
pass  from  the  steam  chest  directly  into  the  exhaust  port,  as  indicated  in  Fig.  35,  or,  as 
the  practical  man  would  say,  the  steam  will  blow  through,  and  therefore  an  unpardon- 
able waste  of  steam  will  be  the  result. 

If  the  outside  edges  of  the  valve  overlap  the  steam  ports,  as  shown  in  Fig.  36,  then 
the  exhaust  cavity  can  be  made  within  certain  limits  a  little  longer,  without  interfering 
with  the  third  condition. 

57.  For  the  sake  of  distinction  we  may  divide  slide-valves  into  two  classes.  In 
one  class  we  may  place  all  slide-valves  whose  outside  edges  just  cover  the  steam  ports, 
such  as  shown  in  Fig.  34.  These  valves  will  admit  steam  into  the  cylinder  during  the 
whole  stroke  of  the  piston,  or,  as  the  practical  man  would  say,  "  the  valve  follows  full 
stroke." 

In  the  other  class  we  may  place  all  slide-valves  whose  outer  edges  overlap  the 
steam  ports,  such  as  shown  in  Fig.  36.  These  valves  will  not  admit  steam  into  the 
cylinder  during  the  whole  stroke  of  the  piston,  but  will  close  the  steam  ports,  and  thus 
cut  off  the  flow  of  the  steam  into  the  cylinder,  before  the  piston  has  reached  the  end  of 
a  stroke.  The  position  of  the  piston  at  the  moment  that  steam  is  cut  off  is  called  the 
point  of  cut-off,  and  this  point  depends  upon  the  amount  of  lap. 

."is.  When  a  valve  of  this  kind  is  placed  in  a  central  position,  that  is,  midway  of  its 
travel,  as  shown  in  Fig.  .'!•>,  then  the  amount  of  overlap  at  each  end  is  called  "outside 


40  MODERN  LOCOMOTIVE    CONSTRUCTION. 

lap,"  or  simply  lap.  Thus,  if  as  in  Fig.  36,  the  valve  overlaps  each  port  |  of  an  inch, 
then  the  valve  is  said  to  have  J  of  an  inch  lap.  Under  no  circumstances  should  one  of 
the  outside  edges  of  the  valve  be  placed  flush  with  an  outside  edge  of  the  steam  port, 
and  then  the  total  amount  of  overlap  at  the  other  end  of  the  valve  be  called  lap,  because 
that  would  be  wrong  according  to  the  universal  acceptation  of  the  term  "  lap." 

TRAVEL   OF   THE   VALVE. 

59.  Since  it  is  always  taken  for  granted  that  the  steam  ports  are  made  just  large 
enough — and  no  more — to  give  a  free  exhaust,  we  must  give  the  valves  that  have  no 
lap,  as  shown  in  Fig.  34,  such  a  travel  that  the  outside  edges  of  these  valves  will  wholly 
open  the  steam  ports.  We  cannot  make  this  travel  any  less,  because  if  it  is  less  the 
steam  port  will  not  be  fully  open  to  the  action  of  the  exhaust.  On  the  other  hand, 
theoretically,  the  travel  of  a  valve  that  has  lap  need  not  be  such  that  the  steam  port 
will  be  fully  opened  to  the  admission  of  steam ;  all  that  is  really  needed  for  this  pur- 
pose is  an  opening  of  ^  of  the  width  of  the  steam  port,  and  if  this  does  not  interfere 
with  the  free  action  of  the  exhaiist,  that  is,  if  it  does  not  prevent  the  full  opening  of 
the  steam  port  for  the  escape  of  steam,  satisfactory  results  will  follow.  In  practice, 
such  niceness  in  the  travel  of  the  valve  is  seldom  aimed  at.  In  fact,  it  is  customary  to 
assign  such  a  travel  to  a  valve  that  the  outside  edges  of  the  valve  will  not  only  fully 
open  the  steam  ports,  but  travel  a  little  beyond  them.  Adopting  such  a  practice,  we 
gain  the  following  advantages :  When  a  valve  that  has  no  lap  travels  a  little  further 
than  necessary  to  fully  open  the  steam  port,  we  have  the  assurance  that  a  slight 
inaccuracy  in  workmanship,  which  cannot  always  be  prevented,  will  not  interfere  with 
the  full  opening  of  the  steam  port.  Again,  it  is  always  desirable  that  when  the  valve 
is  to  cut  off  steam,  it  will  do  so  as  quickly  as  possible,  hence,  when  valves  that  have 
lap  and  their  travel  is  greater  than  absolutely  necessary,  the  motion  of  these  valves 
will  be  quicker  than  the  motion  of  valves  with  shorter  travel ;  therefore,  when  the 
former  are  employed,  the  cut-off  will  be  sharper  and  more  decisive. 

EXAMPLE  13. — To  find  the  travel  of  a  valve  that  has  no  lap,  the  width  of  the  steam 
port  being  given,  let  the  steam  port  be  li  inch  wide.  The  travel  of  a  valve  without 
lap  must  at  least  be  equal  to  twice  the  width  of  the  steam  port.  The  truth  of  this 
must  be  perceived  when  we  examine  Fig.  34,  and  remember  the  remarks  just  made. 
Consequently  the  travel  of  the  valve  in  our  example  will  be  li  x  2  =  2J".  From  the 
foregoing  we  may  establish  the  following : 

RULE  10. — To  find  the  travel  of  a  valve  without  lap,  the  steam  port  to  be  fully 
opened  and  no  more.  Multiply  the  width  of  steam  port  in  inches  by  2,  the  product 
will  be  the  travel  of  the  valve. 

Now,  if  the  travel  of  the  valve  is  to  be  such  that  the  valve  shall  move  i  of  an  inch 
beyond  the  steam  port,  then  we  must  add  this  amount  to  the  width  of  steam  port,  and 
make  the  travel  equal  to  twice  this  sum.  Thus : 

EXAMPLE  14. — Width  of  steam  port  li  inch,  the  valve  to  .travel  i  of  an  inch  beyond 
the  steam  port,  find  the  travel,  li  +  i  ==  li,  and  li  x  2  =  3  inches  =  travel  of  the 
valve.  From  the  foregoing  we  have  the  following : 

RULE  11. — To  find  the  travel  of  the  valve  without  lap,  the  travel  to  extend  a  given 


.MIIHKK\  Loco.tiDT/1-K   ro.v.STV.vrTW.v. 


41 


amount  beyond  the  steam  port  Add  the  given  amount  which  the  valve  must  travel 
beyond  tlic  steam  port  to  the  width  of  the  steam  port,  multiply  the  sum  by  2;  the 
product  will  lie  the  travel  of  the  valve. 

(iO.  When  this  valve  is  used  in  an  engine  with  no  rocker  interposed,  then  the 

eccentricity  of  ill -centric,  or,  in  other  words,  the  distance  between  the  center,  f,  of 

the  crankshaft  and  the  center,  ./•,  of  the  eccentric,  Fig.  30,  will  be  li  inch,  and  the 
throw  of  the  eccentric  will  be  equal  to  the  travel  of  the  valve,  namely,  3  inches. 

To  iii n  1  tin1  travel  of  a  valve  that  has  lap,  the  lap  and  width  of  the  steam  port 
being  given  : 

I'AAMI'LE  1"). — The  width  of  the  steam  port  is  lj  inch,  the  lap  is  J  of  an  inch,  find 
the  travel. 

If  the  valve  is  to  open  the  steam  port  fully,  and  no  more,  for  the  admission  of 
steam,  then  the  travel  cannot  be  less  than  twice  the  sum  of  the  width  of  the  steam  port 
and  lap.  Hence  in  our  example  we  have  1J  +  J  =  2j,  and  2J  x  2  =  4\  inches  =  travel  of 
the  valve.  If  the  valve  is  to  move  J  of  an  inch  beyond  the  steam  port,  then  we  have 
M  +  J  +  I  —  28,  and  2g  x  2  =  4f  inches  =  travel  of  the  valve.  When  no  rocker  is  inter- 
posed, the  throw  of  the  eccentric  is  equal  to  the  travel,  that  is,  4£  inches.  From  the 
foregoing  we  can  establish  the  follow  rules: 

Kn,E  1± — To  find  the  travel  t if  the  valve  with  lap,  the  travel  not  to  extend  beyond 
the  steam  port.  Add  the  width  of  the  steam  port  to  the  lap,  multiply  the  sum  by  2; 
the  product  will  be  the  travel  of  the  valve. 

RULE  13. — To  find  the  travel  of  a  valve  with  lap,  the  travel  to  extend  beyond  the 
steam  port.  Add  the  width  of  the  steam  port,  the  lap,  and  the  amount  of  travel  beyond 
the  steam  port,  multiply  this  sum  by  2;  the  product  will  be  the  travel  of  the  valve. 


POSITION   OF   ECCENTRIC   WHEN   NO   EOCKEK   IS   USED. 

61.  Assume  that  a  valve  without  lap  is  to  be  used  in  an  engine  similar  to  that 
shown  in  Fig.  37,  that  is,  an  engine  in  which  the  axis  of  the  cylinder  will  pass  through 
the  center  of  the  crank-shaft ;  also  let  it  be  required  that  at  the  precise  moment  at 


Backcnd 


/•'if/.  - 


=XJL 


— a 

Front  end 


which  the  piston  readies  the  end  of  a  stroke,  the  valve  shall  open  the  steam  port.  In 
Fig.  37,  it  will  be  noticed,  that  instead  of  showing  the  back  of  the  slide-valve  as  it 
should  be,  we  have  assumed  the  valve  and  seat  to  be  turned  around  the  valve  stem,  SO 
as  to  see  a  section  of  the  valve  ;md  seat  :  this  will  make  the  illustration  more  intelligible 
for  our  purpose.  Now,  assume  that  the  piston  stands  at  I),  the  back  end  of  the  stroke, 
then  the  crank  will  be  in  the  position  as  shown,  and  according  to  our  proposition,  the 


42 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


valve  must  at  that  instant  open  the  back  stearn  port,  consequently  the  valve  must 
stand  in  the  position  as  indicated  in  the  figure.  Again,  assume  that  the  piston  stands 
at  E,  the  front  end  of  the  stroke,  then  the  valve  must  stand  in  a  position  so  that  the 
least  movement  will  open  the  front  steam  port,  consequently  the  valve  will  occupy  the 
same  position  as  before.  In  fact,  when  the  piston  stands  at  either  end  of  the  stroke, 
the  valve,  having  no  lap,  must  stand  in  a  central  position,  that  is,  midway  of  its  travel. 
To  find  the  suitable  position  of  the  eccentric  when  this  valve  stands  central,  we  proceed 
as  follows : 

From  the  center  C  describe  a  circle  a  b  d  (Fig  37),  whose  diameter  is  equal  to 
the  travel  of  the  valve ;  the  circumference  of  this  circle  will  represent  the  path  of  the 
center  of  the  eccentric,  or,  what  amounts  to  the  same  thing,  the  path  of  the  center  of 
the  pin  y ;  consequently  the  exact  location  of  the  center  of  this  pin  must  be  found 
somewhere  in  the  circumference  a  b  d.  Now,  the  diameter  a  b  is  equal  to  the  travel 
of  the  valve,  hence  we  may  assume  that  the  point  a  will  represent  one  end,  and  the 
point  b  the  other  end  of  the  travel,  and  C  the  center  of  the  travel. 

Therefore  to  find  the  location  of  the  pin  y,  draw  through  the  center  C  a  line  i  h 
perpendicular  to  the  line  A  B ;  the  line  i  h  will  intersect  the  circumference  a  I  d  in 
the  points  y  and  d.  If  the  crank-pin  is  to  turn  in  the  direction  of  the  arrow  marked  1, 
then  the  point  y  will  be  the  center  of  the  pin  or  the  center  of  the  eccentric ;  if  the  crank 
is  to  turn  jn  an  opposite  direction,  as  indicated  by  arrow  2,  then  the  point  d  will  be  the 
location  of  the  pin  y. 

From  the  foregoing  we  learn  that  the  center  of  the  eccentric,  or  the  pin  y,  will 
always  travel  ahead  of  the  crank  in  either  direction,  providing  no  rocker  is  used. 
Secondly,  we  learn  that  for  a  valve  without  lap  the  center  of  the  eccentric  will  be  found 
in  a  line  drawn  perpendicular  to  a  line  passing  through  the  center  of  crank-pin  and  the 
center  of  crank-shaft,  providing  the  valve  has  no  lead. 

The  straight  line  drawn  through  the  center  of  crank-shaft  and  center  of  crank-pin 
will,  in  the  future,  be  called  "  the  center  Vine  of  crank." 

62.  By  "  lead "  is  meant  the  width  of  the  opening  of  the  steam  port  at  the  com- 
mencement of  the  stroke  of  the  piston.  Thus,  in  Fig.  38,  when  the  piston  stands  at  the 


beginning  of  the  stroke,  and  the  valve  has  then  opened  the  steam  port  ^  of  an  inch, 
that  rc  of  an  inch  of  opening  is  called  "  lead,"  and  the  valve  is  said  to  have  iV  of  an 
inch  lead.  Again,  if  the  valve  has  opened  the  steam  port  &  of  an  inch,  instead  of  -,-',(  of 
an  inch,  the  valve  is  said  to  have  &  of  an  inch  lead.  In  our  previous  example  the  valve 


l.»r<niiiTirK    COXSTIU  <  ri".\. 


43 


had  no  lf:i<l,  therefore  tin-  <|uestion  arises,  where  shall  we  place  the  eccentric  when  the 
valve  IIMS  li-adf 

K\\Mi'i,K  Iti. — Let  A  l>,  in  Fig.  39,  represent  the  center  line  of  crank;  the  direction 
in  which  the  crank  is  to  move  is  indicated  by  the  arrow.  The  lead  is  to  be  ^  of  an 
inrh,  the  travel  of  the  valve  5  inches;  find  the  position  of  the  eccentric.  From  the 
center  ('  draw  a  circle  a  b  <l,  whose  diameter  is  equal  to  the  travel  of  the  valve, 
namely,  '•>  indies;  on  the  lino  A  li  lay  off  a  point/  /„  of  an  inch  from  the  center  C; 
through  the  point /draw  a  line  k  I  perpendicular  to  the  line  A  li,  this  line  k  I  will 
intersect  the  circumference  a  b  d  in  the  points  y  and  rf;  the  point  y  will  be  the  center 
of  the  eccentric.  If  tin- crank  had  been  designed  to  move  in  the  opposite  direction, 
then  tin-  point  (/would  have  been  the  center  of  the  eccentric.  If  no  lead  had  been 
required,  then  the  center  of  the  eccentric  would  have  been  found  in  the  line  i  h  drawn 
through  the  center  C. 

From  this  we  learn  that  when  the  valve  is  to  have  lead  the  center  of  the  eccentric 
must  lie  moved  forward  of  the  line  *  A,  and  the  amount  that  the  center  of  the  eccentric 
must  be  moved  forward  (or  away  from  the  crank-pin)  is  equal  to  the  lead. 

63.  The  line  i  h  is  an  important  one,  because  the  position  of  the  eccentric  is  always 
laid  off  from  this  line.  In  this  particular  case,  Fig.  39,  the  line  i  h  has  been  drawn 
perpendicular  to  the  center  line  of  crank.  But  from  this  we  must  not  conclude  that  in 
every  case  the  line  i  h  must  be  drawn  perpendicular  to  the  center  line  of  crank.  The 
line  /'  h  in  every  case  is  drawn  perpendicular  to  the  center  line  of  motion  of  the  valve 
gear,  irrespective  of  the  position  of  the  crank.  In  Fig.  39  the  center  line  of  motion  of 
the  valve  gear  coincides  with  the  center  line  of  crank,  and  it  is  for  this  reason  that  the 
line  i  h  has  been  drawn  perpendicular  to  the  latter.  This  will  be  made  plainer  as  we  pro- 
ceed. By  "  the  center  line  of  motion  of  the  valve  gear  "  is  meant  a  line  drawn  through 
the  center  of  the  shaft  parallel  to  the  direction  in  which  the  valve' moves  when  no  rocker 
or  other  mechanism  between  the  shaft  and  valve  is  used.  The  definition  of  the  center 
line  of  the  motion  of  the  valve  gear  in  cases  where  rockers  are  used  will  be  given  later. 


TO  FIND  THE  POSITION   OF  THE  ECCENTRIC  CORRESPONDING  TO  ANY  ONE  OF  THE 
DIFFERENT   POSITIONS   OF  THE  VALVE. 

64.  We  have  stated  in  Art.  55,  that  for  the  construction  of  a  slide-valve  and  for  the 
purpose  of  following  the  movements  of  the  same,  all  that  will  be  required  is  a  section 
0  of  the  valve,  the  valve  seat,  and  the  circumference  of  a  circle 

to  represent  the  path  of  the  center  of  eccentric,  and  such  we 
shall  now  employ  in  the  solution  of  the  following  problems. 

By  the  center  of  ec- 
centric we  mean  the  cen- 
ter x,  Fig.  40,  and  not 
the  center  c  of  the  hole. 

The  center  line  of  MII  ec-  \^_^/  Fiy.4:l 

rnj.40  centric  is  a   straight  line 

drawn  through  the  centers  /  and  ',  and  produced  to  the  circumference  of  the  eccentric 
Fig.  41  shows  the   valve  seat  with  the  steam  ports  and  exhaust  port,  also  the  valve 


44  MODERN  LOCOMOTIVE   CONSTRUCTION. 

standing  in  the  center  of  its  travel.  This  position  of  the  valve  is  an  important  one, 
because  to  this  we  generally  refer  when  any  other  position  of  the  valve  is  to  be  con- 
sidered. From  the  point  c  (the  intersection  of  the  lines  A  B  and  d  c)  as  a  center  and 
a  radius  equal  to  half  the  throw  of  the  eccentric,  a  circle  has  been  drawn ;  the  circum- 
ference of  this  circle  will  represent  the  path  of  the  center  x  of  the  eccentric.  Now,  it 
must  be  remembered  that  in  studying  the  laws  relating  to  the  construction  mid  move- 
ment of  the  slide-valve,  the  length  of  the  eccentric  rod  is  always  considered  to  be  an 
infinite  length,  or,  in  other  words,  that  the  eccentric  acts  precisely  in  the  same  man- 
ner as  a  pin  working  in  a  slotted  cross-head  forged  to  the  valve  stem,  such  as  we  have 
described  in  Arts.  54  and  61.  Keeping  this  in  mind,  the  following  explanation  will  be 
easily  understood : 

Since  the  circumference  a  I  m,  Fig.  41,  represents  the  path  of  the  center  of  the 
eccentric,  it  must  be  readily  perceived  that  the  center  c  of  this  circle  also  represents 
the  center  of  the  shaft,  and  since  the  line  A  B  is  drawn  through  the  center  of  the  shaft, 
we  may  call  it  the  center  line  of  motion  of  the  valve  gear,  because  in  this  particular 
case  the  line  A  B  is  parallel  to  the  direction  of  motion  of  the  valve.  The  diameter 
a  b  of  the  circle  a  l>  m  coincides  with  the  line  of  motion  A  B,  consequently  when  the 
center  of  the  eccentric  is  at  b  the  valve  will  be  at  the  forward  end  of  its  travel,  as 
shown  in  dotted  lines,  and  at  the  same  time  indicating  how  far  the  valve  will  travel 
beyond  the  edge  of  the  steam  port.  When  the  center  of  the  eccentric  is  at  a,  the  valve 
will  then  stand  at  the  opposite  end  of  its  travel,  also  shown  by  dotted  lines.  Now, 
when  the  valve  stands  in  its  central  -position,  the  center  of  the  eccentric  must  also 
stand  in  the  center  of  its  path  from  a  to  b,  and  consequently  will  be  at  m.  To  find  the 
point  m  we  draw  a  straight  line  c  h  through  the  center  c  of  the  circle,  and  perpendicular 
to  the  line  of  motion  A  B,  the  point  of  intersection  of  the  line  c  Ji  and  the  circumference 
will  be  the  point  m.  '  So  also  in  a  similar  manner  we  may  find  the  position  of  the 
center  of  the  eccentric  for  any  other  position  of  the  valve.  For  instance,  if  the  edge 
c  of  the  valve  stands  at  ,9,  we  draw  through  the  point  //  a  line  perpendicular  to  the 
line  A  B ;  the  point  of  intersection  n  of  this  line  and  the  circumference  a  b  m  will  be 
the  center  of  the  eccentric,  when  the  valve  stands  at  g. 

In  order  to  save  time  and  labor,  and  to  make  the  solutions  of  the  problems  as 
simple  as  possible,  it  is  always  best  to  place  the  valve,  as  we  have  done,  in  the  center 
of  its  travel,  and  then  adopt  c,  the  point  of  contact  of  the  outer  edge  of  the  valve 
and  the  valve  seat  as  the  center  of  the  circle  whose  circumference  is  to  represent  the 
path  of  the  center  of  eccentric. 

By  adopting  the  foregoing  suggestion  we  also  gain  the  following  advantages ; 
namely,  we  can  see  at  once  how  far  the  valve  will  travel  beyond  the  steam  port  in 
either  direction;  and  we  can  also  see  how  much  the  exhaust  port  will  be  contracted 
when  the  valve  is  at  end  of  its  travel.  Hence  it  must  be  distinctly  remembered  that 
to  trace  the  motion  of  any  slide-valve,  the  valve  should  lie  placed  in  a  central  position, 
and  the  center  of  circle  whose  circumference  represents  the  path  of  the  center  of  the 
eccentric  should  be  the  point  in  which  one  of  the  outer  edges  of  the  valve  touches 
the  valve  seat. 


\inni:i;.\ 


<-<>.\xn:rrri<>.\ 


45 


I.1M.AK  ADVANCE  OF  THE  VALVE  AND  ANGULAR  ADVANCE  OF  THE  ECCENTHIC. 

(i.">.  In  connection  witli  tin1  setting  of  the  eccentric  two  terms  ai-e  used,  namely,  "lin- 
ear advance  <>t'  tin-  valve"  and  "angular  advance  of  the  eccentric."  Between  these  two 
there  exists  sueh  a  close  relation  that  if  we  change  one  we  must  also  change  the  other, 
and  it'  there  is  no  linear  advance  of  the  valve  there  will  be  no  angular  advance  of  the 
eccentric,  therefore  it  is  of  great  importance  to  understand  the  meaning  of  these  terms. 
In  Fig.  4'2  we  have  shown  the  valve  in  two  positions.  The  dotted  lines  represent 
the  valve  standing  in  the  center  of  its  travel,  and  is  marked  D.  The  section  of  the 
valve  in  t'ull  lines,  marked  E,  represents  its  position  at  the  commencement  of  the 


/.  ; 


/ 


Fig.44 


Fig.45 


stroke  of  the  piston  (it  will  be  noticed  here  that  the  valve  has  lead),  the  distance  from 
c  to  r,  is  called  linear  advance  of  the  valve,  and  is  equal  to  the  lap  and  lead,  plainly 
shown  in  the  figure;  in  short,  linear  advance  of  the  valve  means  the  distance  the  valve 
has  traveled  beyond  its  middle  position  when  the  piston  has  reached  the  end  of  the 
stroke.  When  the  valve  stands  in  the  middle  position  D,  the  center  of  the  eccentric. 
will  be  at  ///,  and  a  line  drawn  from  c  to  m  will  represent  the  center  line  of  the  eccentric. 
Again,  when  the  valve  stands  at  £',  the  center  of  the  eccentric  will  be  at  y,  and  a  line 
drawn  from  c  to  y  will  again  represent  the  center  line  of  the  eccentric.  The  angle 
formed  by  the  lines  c  m  and  c  y  is  called  the  angular  advance  of  the  eccentric;  or,  in 
other  words,  by  angular  advance  is  meant  the  angle  that  is  formed  by  the  position  of 
the  center  line  of  the  eccentric  when  the  piston  is  at  the  commencement  of  the  stroke, 
and  the  position  of  the  center  line  of  the  eccentric  when  the  valve  is  in  the  center  of 
its  travel,  the  length  of  the  eccentric  rod  being  assumed  to  be  infinite. 

tiii.  If  the  lap  is  increased  without  changing  the  travel  and  lead,  as  shown  in  Fig. 
4.'!,  the  linear  advance  will  be  greater,  and  consequently  the  angular  advance  m  c  y  will 
also  lie  increased,  which  is  plainly  indicated  in  the  figure.  Here  also  notice  that  the 
center  r  has  been  moved  away  from  the  outside  edge  of  the  steam  port,  because  when 
the  valve  is  placed  central,  its  outside  edge  r,  will  be  at  r,  which,  according  to  what 

has  been  stated  before,  should  be  adopted  for  the  center  of  the  circle,  whose  circum- 
ference (i  in  l>  represents  the  path  of  the  center  of  the  eccentric.  I>y  so  doini;,  we  see 
at  a  glance  how  far  in  either  direction  the  valve  will  travel. 

In  this  figure  we  notice  that  the  valve  does  not  travel  as  far  beyond  the  inner  edge 
of  the  steam  port  as  in  Fig.  42. 


46  MODERN  LOCOMOTIVE    CONSTRUCTION, 

If  the  lap  is  made  less  than  shown  in  Fig.  42,  without  changing  the  travel  and 
the  lead  of  the  valve,  the  linear  advance  will  be  less,  and  consequently  the  angular 
advance  will  be  smaller.  In  this  particular  case  the  center  c  would  have  to  be  moved 
closer  to  the  edge  of  the  steam  port,  because  the  lap  is  smaller.  When  the  valve  has  no 
lap  and  lead,  as  shown  in  Fig.  44,  there  will  be  no  linear  advance,  and  consequently  no 
angular  advance.  In  this  case  the  center  c  will  be  on  the  outer  edge  of  the  steam  port. 
If  the  travel  of  the  valve  is  the  same  as  that  in  Fig.  43,  and  indicated  by  the  diameter 
of  the  dotted  circle  a  m  b  in  Fig.  44,  we  notice  that  in  the  latter  figure  the  travel  is  too 
great ;  our  circle  shows  that  the  valve  will  travel  beyond  the  bridge,  and  thus  open  the 
exhaust  port  to  the  steam  in  the  steam  chest;  therefore,  the  travel  must  be  reduced,  as 
shown  by  the  circumference  %  ml  I  in  a  full  line.  Lastly,  if  lead  be  given  to  a  valve 
that  has  no  lap,  as  shown  in  Fig.  45,  then  we  have  again  linear  advance,  and  conse- 
quently there  will  be  a  corresponding  amount  of  angular  advance.  In  this  case  the 
center  c  will  remain  on  the  outside  edge  of  the  steam  port. 

TO   FIND   THE   KELATIVE   POSITION   OF   THE   ECCENTKIC   TO   THAT   OF   THE   CRANK. 

67.  In  Art.  61  it  was  shown  how  to  set  an  eccentric  for  a  valve  without  lap  and  lead. 
In  that  particular  example  the  center  line  of  the  eccentric  was  placed  perpendicular  to 
the  center  line  of  crank.  These  relative  positions  are  only  true  for  engines  in  which 
all  the  connections  are  direct,  such  as  shown  in  Fig.  30  (see  Art.  63).  When  other 
connections  are  used,  such  as  rockers,  etc.,  it  may  happen  that  for  some  engines  the 
eccentric  would  have  to  be  set  at  right  angles  to  the  crank,  or,  as  the  practical  man 
would  say,  "  the  eccentric  set  square  with  the  crank,"  and  yet  for  other  engines  this 
would  be  wrong. 

Here  we  will  consider  only  the  relative  position  of  eccentric  to  that  of  the  crank 
in  simple  engines  such  as  represented  by  Fig.  30. 

The  position  of  the  valve  in  Fig.  45  indicates  that  the  piston  stands  at  the  com- 
mencement of  its  stroke,  because  the  small  opening  of  steam  port  there  shown  is  sup- 
posed to  be  lead  and  no  more.  Now,  assume  that  the  point  c  is  not  only  the  center  of 
the  circle  whose  circumference  represents  the  path  of  the  center  of  eccentric,  but  is  also 
the  center  of  the  crank-shaft,  consequently  the  center  line  of  crank  must  pass  through 
c,  and  since  all  the  connections  between  crank-shaft  and  cylinder  are  direct,  the  center 
line  of  crank  must  coincide  with  the  line  A  B.  Again,  the  valve  has  opened  the  left- 
hand  steam  port,  therefore  the  center  P  of  the  crank-pin  must  also  be  on  the  left-hand 
side  of  c,  and  in  the  line  A  B.  Lastly,  the  point  y  is  the  center  of  the  eccentric,  and 
since  the  eccentric  must  travel  ahead  of  the  crank  (Art.  61)  in  this  class  of  engines,  we 
conclude  that  the  crank  is  designed  to  turn  in  the  direction  as  indicated  by  the  arrow. 
In  a  similar  manner,  and  for  similar  reasons,  it  can  be  proved  that  when  the  small 
openings  of  the  steam  ports,  as  shown  in  Figs.  42  and  43,  represent  lead,  then  the 
crank-pin  P  must  occupy  the  position  shown  in  these  figures. 

Notice  now  the  fact  that  in  all  the  Figs.  42,  43,  and  45,  the  center  line  c  m  is  per- 
pendicular to  the  center  line  of  crank  P  c,  and  the  angular  advance  is  laid  off  from  the 
line  c  m  towards  the  right  (away  from  the  crank-pin).  Also  notice  another  important 
fact ;  the  distance  between  the  lines  c  m  and  c:  y  in  all  these  figures  is  equal  to  the 


M<  HI  KI; 


1:  i».\  s  /•/.•  i  'i  •  n<>.\. 


47 


linear  advance;  therefore  in  order  to  find  the  point  y  we  must  lay  off  the  linear  advance 

on  a  line  perpendicular  to  the  line  c  w,  and  not  on  the  arc  tii  //. 

These  facts  are  principles  which  are  applicable  to  every-day  practice.   For  instance : 
(•'A  \MPI.E  17. — Travel  of  the  valve  is  5  inches,  lap  1  inch,  lead  fa  of  an  inch,  and 

the  direction  in  which  the  crank  is  to  move  is  indicated  by  the  arrow,  Fig.  40.     Find 

the  relative  position  of  the  eccentric  to  that  of  the  crank. 

I  >raw  the  straight  line  A  B  as  in  Fig.  46,  let  the  point  P  on  the  line  A  B  represent 

the  center  of  the  crank-pin,  and  the  point  c  on  the  same  line  represent  the  center  of  the 

shaft.     From  c  as  a  center,  and  with  a  radius  equal  to  2£  inches  (which  is  half  the 


Fig.47 


throw  of  the  eccentric),  draw  a  circle  a  b  w,  the  circumference  of  this  circle  will  repre- 
sent the  path  of  the  center  of  the  eccentric.  From  c  lay  off  towards  the  right  a  point 
Cj  Ifa  inch  from  c  (this  I  fa  inch  is  the  sum  of  the  lap  and  lead),  through  c,  draw  a  line 
r,  HI  i  perpendicular  to  A  7?,  this  line  cl  ml  will  intersect  the  circumference  a  b  m  in  the 
point  //,  and  this  point  will  be  the  center  of  the  eccentric  when  the  crank  occupies  the 
position  as  shown.  Through  the  point  c  draw  a  line  perpendicular  to  A  /?,  also  a  line 
through  the  points  c  and  y,  then  the  angle  m  c  y  will  be  the  angular  advance,  and  the 
distance  from  c  to  ^  the  linear  advance. 

These  principles  are  also  applicable  to  shop  practice. 

KXVMPLE  18. — Fig.  40  represents  an  eccentric;  Fig.  47,  a  crank;  Fig.  48,  the  end  of 
a  crank-shaft.  It  is  required  to  fasten  to  the  crank-shaft  the  crank  and  eccentric  in 
their  correct  relative  positions  before  the  shaft  is  placed  in  its  bear- 
ings. The  lap  of  the  valve  is  |  of  an  inch,  and  the  lead  -fa  of  an  inch. 

We  must  first  find  the  half-throw  of  the  eccentric.  In  this  case, 
see  Fig.  40,  the  diameter  of  the  shaft  being  greater  than  the  throw 
of  the  eccentric,  we  must  force  a  strip  of  wood  in  the  hole  of  the  eccen- 
tric, and  on  this  strip  find  the  center  c  of  the  hole,  and  the  center  x 
of  the  eccentric ;  the  distance  between  c  and  x  is  equal  to  half  the  throw.  Through 
c  and  x  draw  a  straight  line  c  g ;  this  line  will  be  recognized  as  the  center  line  of  the 
eccentric. 

Fig.  47. — On  the  face  of  the  crank  draw  through  the  centers  c  and  P  a  straight 
line,  which  will  be  the  center  line  of  crank. 

Fig.  48. — Through  the  center  c  of  the  shaft  draw  any  straight  line,  as  /'_,  /'.. 
From  the  center  r,  and  with  a  radius  equal  to  half  the  throw  of  the  eccentric,  draw  <>n 
the  end  of  the  shaft  a  circle  (/  I/  HI.  On  the  line  /'._,  /'.,  lay  off  a  point  rL,  j ;':  of  an  inch 
from  the  center  c;  through  the  point  t:,  draw  a  straight  line  perpendicular  to  /'_,  /'.,, 


J'.nit  nfaliii/'l 

Fiij.48 


48 


MODERN    LOCOMOTIVE    COXSTllt'CTlOX. 


intersecting  tho  circumference  a  b  m  in  the  point  y;  through  the  points  c  and  i/  draw 
the  straight  line  c  y2. 

Fig.  49. — Place  the  crank  on  the  shaft  so  that  its  center  line  P  c  will  coincide  with 
P2  c  on  the  shaft,  and  then  fasten  the  crank. 

Place  the  eccentric  on  the  shaft  so  that  its  center  line  g  c  will  coincide  with  c  y., 
on  the  shaft,  and  fasten  the  eccentric. 

The  crank  and  eccentric  will  then  have  the  correct  position  on  the  shaft,  and  must 
not  be  changed  for  a  valve  having  &  of  an  inch  lap  and  -fa  of  an  inch  lead.  We  here 


Fiff.49 


F'uj.50 


End  »f  .thrift 
Firj.51 


again  call  the  attention  of  the  reader  to  the  fact  that  these  relative  positions  of  crank 
and  eccentric  are  only  correct  for  engines  in  which  all  the  connections  are  direct,  and 
in  which  no  rocker  is  used. 

When  it  is  necessary  to  place  the  crank  and  the  eccentric  some  distance  from  the 
end  of  the  shaft,  it  will  be  an  easy  matter  to  draw  on  the  outside  of  the  shaft  lines 
through  the  points  P2  and  y.,  parallel  to  the  axis  of  the  shaft,  and  then  set  the  crank 
and  eccentric;  to  these  lines. 

Should  it  so  happen  that  the  throw  of  the  eccentric  is  larger  than  the  diameter  of 
the  shaft  (but  which  seldom  occurs),  draw  on  paper  or  on  a  board  any  straight  line  a  b, 
as  in  Fig.  50,  and  from  a  point  C  on  this  line  as  a  center,  and  with  a  radius  equal  to 
half  the  throw  of  the  eccentric,  draw  the  circle  a  b  m.  Find  the  point  y  in  the  circum- 
ference a  b  m  in  the  same  manner  as  the  point  y  in  Fig.  46  has  been  found,  and  then 
draw  the  lines  m  C  and  y  C.  From  the  center  (7,  and  with  a  radius  equal  to  half  the 
diameter  of  the  shaft,  draw  the  circle  P2  P3,  whose  circumference  will  intersect  the 
straight  lines  m  C  and  y  C  in  the  points  w;)  and  y3.  Through  the  center  c  of  the  shaft, 
see  Fig.  51,  draw  a  straight  line  P2  P3,  and  also  the  line  m.2  c  perpendicular  to  P2  P:J ; 
make  the  arc  m.2  y.2  equal  to  >n3  y.A  in  Fig.  50 ;  draw  on  the  end  of  the  shaft  the  line  y.2  c ; 
set  the  eccentric  to  the  line  y.2  c,  and  the  crank  to  the  line  c  P2. 


THE   EFFECT   OF  LAP. 

68.  When  only  one  slide-valve  is  used  for  the  whole  distribution  of  steam  in  one 
cylinder,  as  in  locomotives,  and  the  valve  has  no  lap,  we  may  justly  name  the  form  of 
such  a  valve  a  primitive  one,  because  valves  without  lap,  or  with  only  a  trifling 
amount,  about  fa  of  an  inch,  were  used  in  locomotives  years  ago,  when  the  great  neces- 
sity for  an  parly  and  liberal  exhaustion  was  not  so  well  understood  as  at  present,  the 
chief  aim  then  being  to  secure  a  timely  and  free  admission  of  steam.  Such  valves,  as 


-:H\  i.i>ct>Mt>TirK  i-(>\sri;rcrii>\.  49 

we  have  stated  before,  will  admit  steam  during  the  whole  length  of  the  stroke,  or,  in 
other  words,  follow  full  stroke,  and  release  the  steam  in  one  end  of  the  cylinder  at  the 
same  moment,  or  nearly  so,  that  the  steam  is  admitted  into  the  other  end;  this  is  cer- 
tainly no  profitable  way  of  using  stearn,  for  the  following  reason: 

The  process  of  exhausting  steam  requires  time,  and  therefore  the  release  of  steam 
should  begin  in  one  end  of  the  cylinder  some  time  before  steam  is  admitted  into  the 
other  end,  or,  we  may  say,  the  steam  which  is  pushing  the  piston  ahead  should  be 
released  before  the  end  of  the  stroke  has  been  reached.  This  cannot  be  accomplished 
with  a  valve  having  no  la}),  and  consequently,  when  such  a  valve  is  used,  there  will  not 
lie  suflicient  time  for  the  exhaustion  of  steam,  thus  causing  considerable  back  pressure 
in  the  cylinder.  In  order  then  to  secure  an  early  exhaust,  lap  was  introduced ;  first,  § 
of  an  inch  lap  was  adopted,  then  £  of  an  inch.  But  it  soon  became  apparent  that 
working  the  steam  expansively  (a  result  of  lap,  besides  gaining  an  early  exhaust)  addi- 
tioiial  economy  in  fuel  was  obtained,  hence  the  lap  was  again  increased  until  it  became 
5  of  an  inch,  and  in  some  cases  1  inch,  and  even  more  than  this.  At  the  present  time 
the  lap  of  a  valve  in  ordinary  locomotives  with  17"  x  24",  or  18"  x  24"  cylinders  is  J  to 
1  inch,  and  in  a  few  cases  slightly  exceeding  this.  From  these  remarks  we  may  justly 
conclude  that  in  these  days  the  purpose  of  giving  lap  to  the  valve  is  to  cause  it  to 
cut  off  steam  at  certain  parts  of  the  stroke  of  the  piston,  so  that  during  the  remaining 
portion  of  the  stroke  the  piston  is  moved  by  the  expansion  of  the  steam.  When  steam 
is  used  in  this  manner,  it  is  said  to  be  used  expansively. 

Now,  since  the  aim  of  giving  lap  to  a  valve,  is  to  cause  it  to  cut  off  steam  at  desig- 
nated parts  of  the  stroke  of  the  piston,  it  will  be  necessary  first  to  study  the  existing 
relation  between  the  motion  of  the  crank-pin  and  the  motion  of  the  piston. 

RELATION  BETWEEN   MOTION  OF  CRANK-PIN  AND  MOTION  OF  PISTON. 

69.  In  order  to  illustrate  this  subject  plainly,  we  have  adopted  in  Fig.  52  a  shorter 
length  for  the  connecting  rod  than  is  used  in  locomotives. 

The  circumference  of  the  circle  A  B  M  I),  drawn  from  the  center  of  the  axle,  and 
with  a  radius  equal  to  the  distance  between  the  center  of  axle  and  that  of  the  crank- 
pin,  represents  the  path  of  the  latter.  We  will  assume  that  the  motion  of  the  crank- 
pin  is  uniform,  that  is,  that  it  will  pass  through  equal  spaces  in  equal  times.  The 
direction  in  which  the  crank-pin  moves  is  indicated  by  the  arrow  marked  1,  and  the 
direction  in  which  the  piston  moves  is  indicated  by  arrow  2. 

In  order  to  trace  the  motion  of  the  piston  it  is  not  necessary  to  show  the  piston  in 
our  illustration,  because  the  connection  between  the  cross-head  pin  P  and  the  piston  is 
rigid ;  hence,  if  we  know  the  motion  of  the  former,  we  also  know  the  motion  of  the 
latter. 

The  line  A  C  represents  the  line  of  motion  of  the  center  of  cross-head  pin  P,  con- 
sequently no  matter  what  position  the  crank  may  occupy,  the  center  P  will  always  be 
found  in  the  line  A.  C.  The  semi-circumference  A  li  I)  will  be  the  path  of  the  center 
of  the  crank-pin  P  during  one  stroke  of  the  piston;  the  point  A  will  be  the  position  of 
the  crank  at  the  beginning  of  the  stroke;  and  //,  the  position  of  the  same  at  the  end  of 
the  stroke.  The  semi-circumference  A  B  D  is  divided  into  12  equal  parts,  although 


50  MODERN  LOCOMQTirE  COXST&VOffOJT. 

any  other  number  would  serve  our  purpose  as  well.  The  distance  between  the  centers 
D  and  P  represents  the  length  of  the  connecting-rod. 

From  the  point  A  as  a  center,  and  with  a  radius  equal  to  D  P  (the  length  of  the 
connecting-rod),  an  arc  has  been  drawn,  cutting  the  line  A  Cin  the  point  a;  this  point 
is  the  position  of  the  center  P  of  the  cross-head  pin,  when  the  center  of  the  crank  is  at 
A.  Again,  from  the  point  1  on  the  semi-circumference  as  a  center  and  with  the  radius 
D  P,  another  arc  has  been  drawn  cutting  the  line  A  C  in  the  point  Ip,  and  this  point 
indicates  the  position  of  the  cross-head  pin  when  the  crank-pin  is  at  the  point  1.  In 
a  similar  manner  the  points  2p,  3p,  4p,  etc.,  have  been  obtained,  and  these  points  indi- 
cate the  various  positions  of  cross-head  pin  when  the  crank-pin  is  in  the  corresponding 
positions,  as  2,  3,  4,  etc. 

Now  notice  the  fact  that  the  spaces  from  A  to  1  and  from  1  to  2,  etc.,  in  the  semi- 
circumference  A  B  D  are  all  equal,  and  the  crank-pin  moves  through  each  of  these 
spaces  in  equal  times,  that  is,  if  it  requires  one  second  to  move  from  A  to  1,  it  will 
also  require  one  second  to  move  from  1  to  2.  The  corresponding  spaces  from  a  to  Ip 
and  from  Ip  to  2p,  etc.,  on  the  line  A  C  are  not  equal,  and  yet  the  cross-head  pin  must 
move  through  these  spaces  in  equal  times ;  if  it  requires  one  second  to  move  from  a  to 
Ip,  it  will  also  require  one  second  to  move  from  Ip  to  1p.  But  this  last  space  is  greater 
than  the  first.  Here,  then,  we  see  that  the  cross-head  pin,  and  therefore  the  piston, 
has  a  variable  motion,  that  is,  the  piston  will,  at  the  commencement  of  its  stroke, 
move  comparatively  slow,  and  increase  in  speed  as  it  approaches  the  center  of  the 
stroke,  and  when  the  piston  is  moving  away  from  the  center  of  stroke,  its  speed  is  con- 
stantly decreasing.  This  variable  motion  of  the  piston  is  mostly  caused  by  changing 
a  rectilinear  motion  into  a  uniform  rotary  motion,  and  partly  by  the  angle  formed  by 
the  center  line  D  P  of  the  connecting-rod  and  the  line  A  C,  an  angle  which  is  con- 
stantly changing  during  the  stroke.  Also  notice  that  the  distance  from  a  to  Ip  nearest 
one  end  of  the  stroke  is  smaller  than  the  distance  from  b  to  lip  nearest  the  other  end 
of  the  stroke,  and  if  we  compare  the  next  space  Ip  to  2p  with  the  space  Up  to  IQp,  we 
again  find  that  the  former  is  smaller  than  the  latter,  and  by  further  comparison  we  find 
that  all  the  spaces  from  a  to  6p  are  smaller  than  the  corresponding  spaces  from  b  to  6p, 
and  consequently  when  the  crank-pin  is  at  point  6,  which  is  the  center  of  the  path 
of  the  crank-pin  during  one  stroke,  the  cross-head  pin  P  will  be  at  6p,  and  not  in  the 
center  of  its  sti'oke.  Thus  we  see  that  the  motion  of  the  piston  is  not  symmetrical,  and 
this  is  wholly  due  to  the  varying  angularity  of  the  connecting-rod  during  the  stroke. 
If  we  make  the  connecting-rod  longer,  but  leave  the  stroke  the  same,  the  difference 
between  the  spaces  b  to  lip  and  a  to  Ip  will  be  less,  and  the  same  can  be  said  of  the 
other  spaces.  Again,  if  we  consider  the  length  of  the  connecting-rod  to  be  infinite, 
then  the  difference  between  the  spaces  nearest  the  ends  of  the  stroke  will  vanish,  and 
the  same  result  is  true  for  the  other  spaces.  Hence,  when  the  length  of  the  connect- 
ing-rod is  assumed  to  be  infinite,  the  motion  of  the  piston  will  be  symmetrical,  but 
still  remain  variable ;  in  fact,  the  piston  will  haveithe  same  motion  as  that  shown  in  Fig. 
53.  In  this  figure  we  have  dispensed  with  the  connecting-rod,  and  in  its  place  extended 
the  piston  rod,  and  attached  to  its  end  a  slotted  cross-head  in  which  the  crank-pin  is 
to  work.  Although  such  mechanism  is  never  used  in  a  locomotive,  yet  with  its  aid  we 
can  establish  a  simple  method  for  finding  the  position  of  the  piston  when  that  of  the 


MODKK.\   LOCOMUl'lfK   (OXtiTRUCTION. 


51 


crank  is  known.  In  this  figure,  as  in  Fig.  52,  the  circumference  A  B  D  M  will  repre- 
sent the  path  of  the  center  of  the  crank-pin,  and  from  the  nature  of  this  mechanism  it 
must  be  evident  that  at  whatever  point  in  the  circumference  A  B  D  M  the  crank-pin 


center  may  lie  located,  the  center  line  i  h  of  the  slotted  cross-head  will  always  stand 
perpendicular  to  the  line  A  6',  and  also  pass  through  the  center  of  crank-pin. 

In  Fig.  5:!,  when  the  crank-pin  is  at  A,  the  piston  will  be  at  the  commencement  of 
its  stroke.  During  the  time  the  crank-pin  travels  from  A  to  point  8  the  piston  will 
travel  through  a  portion  of  its  stroke  equal  to  the  length  A  E,  which  is  the  distance 
between  the  dotted  line  i  h  and  the  full  line  i  h.  If  now  we  assume  the  points  1,  2,  3, 
etc.,  in  the  semi-circumference  A  B  I)  to  be  the  various  positions  of  the  crank-pin  dur- 
ing one  stroke,  and  then  draw  through  these  points  lines  perpendicular  to  the  line  A 
C,  cutting  the  latter  in  the  points  1pt  2p,  3p,  etc.,  we  obtain  corresponding  points  for  the 


fig.  54 


position  of  the  piston  in  the  cylinder.  Thus,  for  instance,  when  the  crank-pin  is  at 
point  1,  the  piston  will  then  have  moved  from  the  commencement  of  its  stroke  through 
a  distance  equal  to  A  lj>,  and  when  the  crank-pin  is  at  point  2,  the  piston  will  then 
have  traveled  from  A  to  '2/>,  and  so  on. 

70.  From  the  fort-going,  wt-  can  establish  a  simple  method,  as  shown  in  Fig.  54, 
for  finding  the  position  of  the  piston  when  that  of  the  crank  is  known.  The  diameter 
A  II  re]. resents  the  stroke  of  the  piston,  and  the  semi-circumference  A  B  1)  represents 
the  path  of  the  center  of  the  crank-pin  during  one  stroke.  For  convenience,  we  may 


52 


MODEHN  LOCOMOTIVE    CONSTRUCTION. 


divide  the  diameter  into  an  equal  number  of  parts,  each  division  indicating  one  inch  of 
the  stroke.  In  this  particular  case  (Fig.  54),  we  have  assumed  the  stroke  to  be  24 
inches ;  hence  the  diameter  has  been  divided  into  24  equal  parts.  Let  the  arrow  indi- 
cate the  direction  in  which  the  crank  is  to  turn,  and  A  the  beginning  of  the  stroke ; 
then,  to  find  the  distance  through  which  the  piston  must  travel  from  the  commence- 
ment of  its  stroke  during  the  time, that  the  crank  travels  from  A  to  b,  we  simply  draw 
through  the  point  b  a  straight  line  7;  c  perpendicular  to  A  B  ;  the  distance  between  the 
line  b  c  and  the  point  A  will  be  that  portion  of  the  stroke  through  which  the  piston  has 
traveled,  when  crank-pin  has  reached  the  point  b.  In  our  figure  we  notice  that  the  line 
b  c  intersects  A  B  in  the  point  6 ;  hence  the  piston  has  traveled  six  inches  from  the 
commencement  of  the  stroke. 

If  this  method  of  finding  the  position  of  the  piston  when  that  of  the  crank  is 
known  is  thoroughly  understood,  then  the  solutions  of  the  following  problems  relating 
to  lap  of  the  slide  valve  will  be  comparatively  easy. 


PBOBLEMS   RELATING   TO   LAP   OF   THE   SLIDE-VALVE. 

71.  To  find  the  point  of  cut-off  when  the  lap  and  travel  of  the  valve  are  given,  the 
valve  to  have  no  lead. 

The  principles  upon  which  the  following  problems  relating  to  the  construction  of 
the  slide-valve  are  based,  have  been  taken  from  the  excellent  "  Practical  Treatise  on 
the  Movement  of  Slide- Valves  by  Eccentrics,"  by  Prof.  C  W.  MacCord. 

EXAMPLE  19. — Lap  of  valve  is  one  inch;  travel,  5  inches;  no  lead;  stroke  of 
piston,  24  inches.  At  what  part  of  the  stroke  will  the  steam  be  cut  off  ? 

We  must  first  find  the  center  c,  Fig.  55,  of  the  circle  a  b  m,  whose  circumference 
represents  the  path  of  the  center  of  eccentric,  and  this  is  found,  as  the  reader  will  re- 

]  member,  by  placing  the  valve 

in  a  central  position  (Art.  64), 
as  shown  in  dotted  lines  in 
this  figure.  Then  the  edge  c 
of  the  valve  will  be  the  center 
of  the  circle.  The  valve  drawn 
in  full  lines  shows  its  position 


at  the  commencement  of  the 
stroke  of  piston  ;  and  since 
the  valve  is  to  have  no  lead, 
the  edge  Cz  will  coincide  with 
the  outer  edge  of  the  steam 
port.  Through  the  edge  €2  draw  the  line  i  h  perpendicular  to  the  line  A  B ;  the  line 
i  h  will  intersect  the  circumference  a  b  m  in  the  point  y,  and  this  point  will  be  the 
center  of  eccentric  when  the  piston  is  at  the  beginning  of  its  stroke.  Now,  assume 
that  the  circumference  a  b  m  also  represents,  on  a  small  scale,  the  path  of  the  center 
of  the  crank-pin ;  then  the  diameter  y  x  of  this  circle  will  represent  the  length  of  the 
stroke  of  the  piston ;  the  position  of  this  diameter  is  found  by  drawing  a  straight  line 
through  the  point  y  (the  center  of  the  eccentric  when  the  piston  is  at  one  end  of  its 


MODERN  LOCOMOTITE   CONSTRUCTION.  53 

stroke)  and  the  renter  r.  Also  assume  that  the  pointy  represents  the  center  of  the 
crank-pin  when  the  piston  is  at  the  beginning  of  its  stroke.  To  make  the  construction 
as  plain  as  possible,  divide  the  diameter  y  x  into  24  equal  parts,  each  representing  one 
inch  of  the  stroke  of  piston,  and  for  convenience  number  the  divisions  as  shown.  The 
arrow  marked  1  shows  the  direction  in  which  the  valve  must  travel,  and  arrow  2  indi- 
cates the  direction  in  which  the  center  y  must  travel.  Now  it  must  be  evident,  because 
the  points  >i  and  C-i  will  always  be  in  the  same  line,  that  during  the  time  the  center  y 
of  the  eccentric  travels  through  the  arc  y  g,  the  valve  not  only  opens  the  steam  port, 
but,  as  the  circumference  a  b  m  indicates,  travels  a  little  beyond  the  port,  and  then 
closes  the  same,  or,  in  short,  during  the  time  the  center  of  eccentric  travels  from  y  to 
//,  the  port  has  been  fully  opened  and  closed;  and  the  moment  that  the  center  of  eccen- 
tric readies  the  point  g,  the  admission  of  steam  into  the  cylinder  is  stopped.  We  have 
assumed  that  the  point  y  also  represents  the  position  of  the  center  of  crank-pin  at  the 
beginning  of  the  stroke ;  and,  since  the  crank  and  eccentric  are  fastened  to  the  same 
shaft,  it  follows  that  during  the  time  the  center  of  eccentric  travels  from  y  to  g  the 
crank-pin  will  move  through  the  same  arc,  and  when  the  steam  is  cut  off  the  crank-pin 
will  be  at  the  point  g.  Therefore,  through  the  point  g  draw  a  straight  line  g  k  per- 
pendicular to  the  line  y  x ;  the  line  g  k  will  intersect  the  line  y  x  in  the  point  k,  and 
this  point  coincides  with  the  point  mark  20 ;  hence  steam  will  be  cut  off  when  the  piston 
has  traveled  I'd  inches  from  the  beginning  of  its  stroke. 

The  manner  of  finding  the  point  k  is  precisely  similar  to  that  of  finding  the  point 
c,  in  Fig.  54.  The  angle  m  y  c  will  be  the  angular  advance  of  the  eccentric. 

LEAD  WILL  AFFECT  THE  POINT  OF  CUT-OFF. 

72.  In  Fig.  f>5  the  valve  had  no  lead ;  if,  now,  in  that  figure,  we  change  the  angular 
advance  m  </  c  of  the  eccentric  so  that  the  valve  will  have  lead,  as  shown  in  Fig.  56, 
then  the  point  of  cut-off  will  also  be  changed.  How  to  find  the  point  of  cut-off  when 
the  valve  has  lead  is  shown  in  Fig.  56. 

EXAMPI.K  '10. — The  lap  of  valve  is  1  inch,  its  travel  5  inches ;  lead  J  of  an  inch  (this 
large  amount  of  lead  has  been  chosen  for  the  sake  of  clearness  in  the  figure);  stroke 
of  piston,  -J4  indies  ;  at  what  part  of  the  stroke 
will  the  steam  be  cut  off? 

On  the  line  A  77,  Fig.  f>(>,  lay  off  the  exhaust 
and  steam  ports ;  also  on  this  line  find  the  cen- 
ter  i-  of  the  circle  <i  //  ///  in  a  manner  similar 
to  that  followed  in  the  last  construction, 
namely,  by  placing  the  valve  in  a  central  »^= 

position,  as  shown  by  the  dotted   lines  and 

marked  I),  and  then  adopting  the  edge  c  of  the  valve  as  the  center  of  the  circle 
a  li  in;  or,  to  use  fewer  words,  we  may  say  from  the  outside  of  the  edge  s  of  the 
steam  port,  lay  off  on  the  line  .1  //a  point  <•  whose  distance  from  the  edge  s  will  be 
equal  to  the  lap,  that  is,  1  inch.  From  '•  as  a  center,  and  with  a  radius  of  24  inches 
(equal  4  of  the  travel),  describe  the  circle  a  b  m,  whose  circumference  will  represent  the 
path  of  the  center  of  eccentric.  The  lead  of  the  valve  in  a  locomotive  is  generally  :;'.., 


54 


MODERN  LOCOMOTIVE   COXSTKUCTIOX. 


and  sometimes  as  much  as  y-8-  of  an  inch,  when  the  valve  is  in  full  gear,  in  this  exam- 
ple we  have  adopted  a  lead  of  J  of  an  inch  for  full  gear,  hence,  draw  the  section 
of  the  valve,  as  shown  in  full  lines,  in  a  position  that  it  will  occupy  when  the  pis- 
ton is  at  the  beginning  of  its  stroke,  and  consequently  the  distance  between  the  edge 
c.,  of  the  valve  and  the  edge  s  of  the  steam  port  will,  in  this  case,  be  J  of  an  inch. 
Through  r2  draw  a  straight  line  perpendicular  to  A  B,  intersecting  the  circumference  a 
b  m  in  the  point  y ;  this  point  will  be  the  center  of  the  eccentric  when  the  piston  is  at 
the  beginning  of  its  stroke,  and  since  it  is  assumed  that  the  circumference  a  b  m  also 
represents  the  path  of  the  center  of  the  crank-pin,  the  point  y  will  also  be  the  position 
of  the  center  of  the  crank-pin  when  the  piston  is  at  the  commencement  of  its  stroke. 
Through  the  points  y  and  c  draw  a  straight  line  y  x,  to  represent  the  stroke  of  the  pis- 
ton, and  divide  it  into  24  equal  parts.  Through  the  point  s  draw  a  straight  line  per- 
pendicular to  A  B,  intersecting  the  circumference  a  b  m  in  the  point  </,  and  through  <y 
draw  a  straight  line  perpendicular  to  y  x,  and  intersecting  the  latter  in  the  point  k ; 
this  point  will  be  the  point  of  cut-off.  If  now  the  distance  between  the  point  k  and 
point  19  is  £  of  the  space  from  19  to  20,  we  conclude  that  the  piston  has  traveled  19J 
inches  from  the  beginning  of  its  stroke  when  the  admission  of  steam  into  the  cylinder 
is  suppressed. 

Here  we  see  that  when  a  valve  has  no  lead,  as  in  Fig.  55,  the  admission  of  steam 
into  the  cylinder  will  cease  when  the  piston  has  traveled  20  inches;  and  when  the 
angular  advance  of  the  eccentric  is  changed,  as  in  Fig.  56,  so  that  the  valve  has  \  of  an 
inch  lead,  the  point  of  cut-off  will  be  at  19J  inches  from  the  beginning  of  the  stroke,  a 
difference  of  £  of  an  inch  between  the  point  of  cut-off  in  Fig.  55  and  that  in  Fig.  56. 
But  the  lead  in  locomotive  valves  in  full  gear  is  only  about  -<fa  of  an  inch,  which  will 
affect  the  point  of  cut-off  so  very  little  that  we  need  not  notice  its  effect  upon  the 
period  of  admission,  and,  therefore,  lead  will  not  be  taken  into  consideration  in  the 
following  examples. 


THE   TRAVEL   OF   THE   VALVE   WILL   AFFECT   THE   POINT   OF   CUT-OFF. 

73.  Fig.  57  represents  the  same  valve  and  ports  as  shown  in  Fig.  55,  but  the  travel 
of  the  valve  in  Fig.  57  has  been  increased  to  5f  inches.     The  point  of  cut-off  k  has  been 

obtained  by  the  same  method  as  that  employed 
in  Figs.  55  and  56,  and  we  find  that  this  point 
k  coincides  with  point  21.  Now  notice  the 
change  caused  by  an  increase  of  travel ;  when 
the  travel  of  the  valve  is  5  inches,  as  shown 
in  Fig.  55,  the  admission  of  steam  into  the 
cylinder  will  cease  when  the  piston  has  trav- 
eled 20  inches  from  the  commencement  of  its 
stroke,  and  when  the  travel  of  the  same  valve 
is  increased  f  of  an  inch,  as  shown  in  Fig.  57,  the  admission  of  the  steam  will  not  be 
suppressed  until  the  piston  has  traveled  21  inches.  Here  we  notice  a  difference  of 
1  inch  between  the  two  points  of  cut-off.  Bi;t  it  must  be  remembered  that  when  the 
travel  of  a  valve  for  a  new  engine  is  to  be  found  or  established,  the  point  of  cut-off 


MODEKN  LOCOMOTIVE   CONSTRUCTION. 


does  not  enter  the  question;  we  simply  assign  such  a  travel  to  the  valve  that  steam 
ports  will  be  fully  opened,  or  give  it  a  slightly  greater  travel  when  the  valve  is  in  full 
gear;  and  how  to  find  this  travel  has  been  explained  in  Art.  59.  The  point  of  cut-off 
is  regulated  by  the  lap  and  position  of  the  eccentric. 

74.  In  order  to  find  the  point  of  cut-off  it  is  not  necessary  to  make  a  drawing  of  the 
valve,  as  has  been  done  in  Fig.  55.  The  only  reason  for  doing  so  was  to  present  the 
method  of  finding  the  point  of  cut-off  to  the  beginner  in  as  plain  a  manner  as  possible. 
In  order  to  show  how  such  problems  can  be  solved  without  the  section  of  a  valve,  and 
consequently  with  less  labor,  another  example,  similar  to  Example  19,  is  introduced. 

I'A\Mi'i.K  21. — Lap  of  valve  is  If  inches;  travel,  5£  inches;  stroke  of  piston,  24 
inches:  width  of  steam  port,  1  \  inches;  find  the  point  of  cut-off. 

Fig.  .~>S.  Draw  any  straight  line,  as  A  B;  anywhere  on  this  line  mark  off  1J  inches, 
equal  to  the  width  of  the  steam  port.  From  the  edge  s  of  the  steam  port  lay  off  on 
the  line  A  B  &  point  c,  the  distance  between  the  points  s  and  c  being  If  inches,  that  is. 


equal  to  the  amount  of  lap.  From  c  as  a  center,  and  with  a  radius  equal  to  half  the 
travel,  namely,  2if  inches,  draw  a  circle  a  b  m;  the  circumference  of  this  circle  will  rep- 
resent the  path  of  the  center  of  the  eccentric,  and  also  that  of  the  (-rank-pin.  Through 
.s  draw  a  straight  line  i  h  perpendicular  to  A  Ji\  this  line  i  li  will  intersect  the  circum- 
feivnee  H  li  m  \\\  the  points  //  and  </.  Through  the  points  y  and  c  draw  a  straight  line 
y  x;  the  diameter  //  ./•  will  represent  the  stroke  of  the  piston.  Divide  y  x  into  24  equal 
parts;  through  the  point  //  draw  a  straight  line  y  k  perpendicular  to  //  .r,  and  intersect- 
ing //  .r  in  the  point  £,  this  point  is  the  point  of  cut-off.  Since  k  coincides  with  the 
point  18,  it  follows  that  the  piston  had  traveled  IS  inches  from  the  beginning  of  its 
stroke  when  the  flow  of  the  steam  into  the  cylinder  ceased. 

7").  Now  we  may  reverse  the  order  of  this  construction  and  thus  find  the  amount 
of  lap  required  to  cut  off  steam  at  a  given  portion  of  the  stroke. 

EXAMPLE  22. — Travel  of  valve  is  .Vf  inches;  stroke  of  piston,  30  inches;  steam  to 
be  cut  off  when  the  piston  lias  traveled  22  inches  from  the  beginning  of  the  stroke; 
find  the  lap. 

Fig.  59.  Draw  a  circle  a  l>  in  whose  diameter  is  equal  to  the  travel  of  the  valve, 
viz.,  5J  inches.  Through  the  center  r  draw  the  diameter // r.  In  this  figure  we  have 
drawn  the  line  //  x  vertically,  which  was  done  for  the  sake  of  convenience:  any  other 
position  for  this  line  will  answer  the  purpose  equally  well.  The  circumference  a  h  m 


56 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


represents  the  path  of  the  center  of  the  eccentric,  also  that  of  the  crank-pin ;  the  diame- 
ter y  x  will  represent  the  stroke  of  the  piston,  and  therefore  is  divided  into  30  equal 
parts.  The  steam  is  to  be  cut  off  when  the  piston  has  traveled  22  inches  from  the 
beginning  of  the  stroke,  therefore  through  the  point  22  draw  a  straight  line  g  k  perpen- 
dicular to  y  x,  the  line  g  k  intersecting  the  circumference  a  b  m  in  the  point  g.  Join 
the  points  y  and  g  by  a  straight  line.  Find  the  center  s  of  the  line  y  g,  then,  through  s 
and  perpendicular  to  the  line  y  g,  draw  the  line  A  B ;  if  the  latter  line  is  drawn  accu- 
rately it  will  always  pass  through  the  center  c.  The  distance  between  the  points  s  and 
c  will  be  the  amount  of  lap  required,  and  in  this  example  it  is  1-^  inches. 

Examples  like  the  foregoing  are  often  given  in  a  somewhat  different  form.  For 
instance,  let  the  travel  of  the  valve  be  5f  inches,  stroke  30  inches,  steam  to  be  cut  off 
at  |  stroke ;  find  the  lap. 

Here  we  draw  the  circle  a  fe  m  and  the  diameter  y  x  as  before ;  but  instead  of  divid- 
ing the  diameter  x  y  into  30  equal  parts  to  correspond  to  number  of  inches  in  the 
stroke,  we  divide  it  into  four  equal  parts ;  the  point  of  cut-off  k  will  then  be  at  f  of  the 
diameter  from  its  extremity  y.  Through  the  point  k  draw  k  g  perpendicular  to  y  x,  and 
proceed  as  before,  and  thus  obtain  the  lap  required. 

It  may  also  be  stated  that  this  construction  will  give  the  amount  of  opening  of 
the  steam  poi't ;  thus,  in  Fig.  59  the  distance  from  s  to  6  shows  the  amount  of  opening 
of  the  steam  port.  If,  for  instance,  s  b  is  equal  to  the  width  of  the  steam  port,  the 
latter  will  be  opened  fully;  if  s  I  is  greater  than  the  width  of  the  steam  port,  the 
edge  of  the  valve  will  travel  beyond  the  inner  edge  of  the  gteam  port ;  and  if  s  b  is  less 
than  the  width  of  the  steam  port,  the  latter  will  not  be  opened  fully.  This  is  obvious 
from  what  has  been  said  in  relation  to  Figs.  33,  41,  42,  43,  and  44. 

76.  It  sometimes  occurs  in  designing  a  new  locomotive,  and  often  in  designing 
stationary  or  marine  engines,  that  only  the  width  of  steam  port  and  point  of  cut-off  is 

known,  and  the  lap  and  travel  of  the  valve  is  not  known. 
In  such  cases  both  of  these  can  be  at  once  determined  by 
the  following  method. 

EXAMPLE  23. — The  width  of  the  steam  port  is  2  inches ; 
the  stroke  of  piston,  30  inches ;  steam  to  be  cut  off  when 
the  piston  has  traveled  24  inches  from  the  beginning  of  its 
stroke ;  find  the  lap  and  travel  of  the  valve. 

Fig.  60.  Draw  any  circle,  as  A  B  M,  whose  diameter  is 
larger  than  what  the  travel  of  the  valve  is  expected  to  be. 
Through  the  center  c  draw  the  diameter  y  x,  and,  since  the 
stroke  of  the  piston  is  30  inches,  divide  y  x  into  30  equal 
parts.  Steam  is  to  be  cut  off  when  the  piston  has  traveled  24  inches ;  therefore  through 
point  24  draw  a  straight  line  g  k  perpendicular  to  the  diameter  y  x,  intersecting  the  cir- 
cumference ABM  in  the  point  .</.  Join  the  points  y  and  g  by  a  straight  line ;  through 
the  center  s  of  the  line  y  g  draw  a  line  A  B  perpendicular  to  y  g.  So  far,  this  construction 
is  precisely  similar  to  that  shown  in  Fig.  59,  and  in  order  to  distinguish  this  part  of  the 
construction  from  that  which  is  to  follow,  we  have  used  dotted  lines ;  for  the  remainder 
full  lines  will  be  used.  It  will  also  be  noticed  by  comparing  Fig.  60  with  Fig.  59  that, 
if  the  diameter  A  B  had  been  the  correct  travel  of  valve,  then  c  s  would  have  been  the 


MOI/I-:I;\  i.ui  IIMIITI (•/•;  COXSTRUCTIOX.  57 

correct  amount  of  lap.  But  we  commenced  this  construction  with  a  travel  that  we 
knew  to  be  too  great ;  hence,  to  find  the  correct  travel  and  lap,  we  must  proceed  as 
follows :  Join  the  points  />'  and  y.  From  s  towards  li,  lay  off  on  the  line  A  B  a  point 
b ;  the  distance  between  the  points  s  and  b  must  be  equal  to  the  width  of  the  steam 
port  plus  the  amount  that  the  valve  is  to  travel  beyond  the  steam  port,  which,  in  this 
example,  is  assumed  to  be  £  of  an  inch.  Therefore  the  distance  from  s  to  b  must  be 
'2k  inches.  Through  b  draw  a  straight  line  b  y.2  parallel  to  B  y,  intersecting  the  line  y 
ff  in  the  point  //._>.  Through  the  point  v/2  draw  a  straight  line  y.2  #2  parallel  to  the  line 
// ./,  and  intersecting  the  line  A  B  in  the  point  c.2.  From  c.>  as  a  center,  and  with  a 
radius  equal  to  c2  fc,  or  c2  «/2,  describe  a  circle  a  b  y.,.  Then  a  b  will  be  the  travel  of  the 
valve,  which,  in  this  case,  is  7§  inches,  and  the  distance  from  c2  to  s  will  be  the  lap, 
which,  in  this  example,  is  Ij-J  inches. 

PRACTICAL  CONSTRUCTION   OF  THE   SLIDE-VALVE. 

77.  It  should  be  obvious,  and  therefore  almost  needless  to  remark  here,  that  the 
foregoing  graphical  methods  employed  in  the  solutions  of  the  problems  relating  to  the 
slide-valve  are  applicable  to  every-day  practice.  The  writer  believes  that  these  methods 
are  the  simplest  and  best  to  adopt  for  ordinary  use,  and  without  these  it  would  be  diffi- 
cult to  construct  a  valve  capable  of  performing  the  duty  assigned  to  it.  Of  course, 
when  a  graphical  method  is  employed,  great  accuracy  in  drawing  the  lines  is  necessary. 

We  will  give  a  practical  example,  in  which  one  of  the  objects  aimed  at  is  to  show 
the  application  of  one  of  the  foregoing  methods  to  ordinary  practice. 

EXAMPLE  24. — The  width  of  the  steam  ports  is  1|  inches ;  length  of  the  same,  14 
inches;  thickness  of  bridges,  lj  inches;  width  of  exhaust  port,  2£  inches;  travel  of 
valve,  4f  inches;  stroke  of  piston,  24  inches;  steam  to  be  cut  off  when  the  piston 
has  traveled  20f  inches  from  the  beginning  of  its  stroke;  the  edges  of  the  exhaust 
cavity  are  to  cover  the  steam  ports,  and  not  more,  when  the  valve  stands  in  a  central 
position ;  construct  the  valve. 

Fig.  61.  Draw  a  straight  line  A  Bio  represent  the  valve  seat ;  through  any  point  in 
A  B  draw  another  line  D  C  perpendicular  to  A  B ;  the  line  D  C  is  to  represent  the 
center  of  exhaust  port  and  the  center  of  valve.  Draw  the 
exhaust  port,  bridges,  and  steam  ports  as  shown. 

The  question  now  arises :  How  long  shall  we  make  the 
valve?  Or,  in  other  words,  what  shall  be  the  distance  be- 
tween the  outside  edges  of  the  valve  <•  and  c2  ?  If  the  valve  had 
to  admit  steam  during  the  whole  stroke  of  the  piston,  or,  as  the  <r  * '"' Cj 

practical  man  would  say,  "  follow  full  stroke,"  then  the  distance  between  the  edges  rand 
<-.,  would  be  equal  to  the  sum  of  twice  the  width  of  one  steam  port  pins  twice  the  width 
of  one  bridge  plus  the  width  of  the  exhaust  port,  hence  we  would  have  2£  +  2J  +  2J  =  7J 
inches  for  the  length  of  the  valve.  But  according  to  the  conditions  given  in  the  exam- 
ple, the  valve  must  cut  off  steam  when  the  piston  has  traveled  20?  indies,  therefore 
the  valve  must  have  lap,  and  the  amount  of  lap  that  is  necessary  for  this  purpose  must 
be  determined  by  the  method  shown  in  Fig.  .">!),  and  given  in  connection  with  Kxample 
22.  Following  this  method,  we  find  that  the  required  lap  is  \  of  an  inch,  therefore  the 


58 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


total  length  of  the  valve  will  be  7i  +  (i  x  2)  =  9  inches :  or  we  may  say  that  the  distance 
between  the  edges  c  and  c2  must  be  equal  to  twice  the  width  of  one  steam  port  plus 
twice  the  width  of  one  bridge  plus  the  width  of  the  exhaust  port  plus  twice  the  lap, 
consequently  we  have  2£+2j+2£+l£  =  9  inches  for  the  length  of  the  valve.  Through 
the  points  c  and  c2  (each  point  being  placed  4£  inches  from  the  center  line  C  />),  draw 
lines  perpendicular  to  A  B ;  these  lines  will  represent  the  outside  surfaces  containing 
the  edges  c  and  c2.  These  surfaces  must  be  square  with  the  surface  A  B,  because,  if 
they  are  not  so,  but  are  such  as  shown  in  Fig.  64,  the  distance  between  the  edges  c  and 
c.2  will  decrease  as  the  valve  wears,  and  when  this  occurs  the  valve  will  not  cut  off  the 
steam  at  the  proper  time.  Now,  in  regard  to  the  cavity  of  the  valve.  One  of  the  con- 
ditions given  in  our  example  is,  that  the  edges  of  the  cavity  must  cover  the  steam 
ports,  and  no  more,  when  the  valve  stands  in  a  central  position,  therefore  the  inner 
edges  i  and  i2  of  the  valve  must  be  4g  inches  apart,  which  is  equal  to  twice  the  width 
of  one  bridge  plus  the  width  of  the  exhaust  port ;  consequently,  when  the  valve  stands 
midway  of  its  travel,  the  inner  edges  of  the  valve  (being  4$  inches  apart)  and  the  inner 
edges  of  the  steam  ports  coincide.  Through  the  points  i  and  i,  (each  being  placed  2§ 
inches  from  the  center  line  C  D),  draw  the  straight  lines  i  e  and  i.2  c.2  perpendicular  to 
A  B.  These  lines  will  represent  the  sides  of  the  cavity  containing  the  inner  edges  * 
and  i2  of  the  valve,  and  these  sides  must  be  square  with  the  surface  A  B ;  if  these  are 
otherwise,  for  instance  such  as  shown  in  Tig.  64,  the  distance  between  the  edges  i  and 
i.2  will  change  as  the  valve  wears,  and  then  the  valve  will  not  perform  its  duty  correctly. 
The  depth  d  d2  of  the  cavity  is  generally  made  from  1^  to  l£  times  the  width  of  the 
exhaust  port.  The  writer  believes  that  making  the  depth  of  the  cavity  l£  times  the 


1'iy.  C-t 


width  of  the  exhaust  port  is  the  best  practice.  In  our  example  the  width  of  the  exhaust 
port  is  2£  inches,  and  2£  x  1J  =  3f  inches,  which  will  be  the  distance  from  d  to  d.2,  that 
is,  the  depth  of  the  cavity.  The  curved  surface  of  the  cavity  is  generally  a  cylindrical 
surface,  and  when  it  is  so,  as  in  our  example,  this  surface  must  be  represented  in  Fig. 
61  by  an  arc  of  a  circle.  The  sides  i  e  and  i2  e.2  must  be  planed,  and  to  do  this  conven- 
iently, these  sides  must  extend  a  little  beyond  the  curved  surface,  towards  the  center 
C  D.  Consequently,  through  the  point  d.2  draw  an  arc  whose  center  is  in  the  line  C  Z), 
and  whose  radius  is  such  that  will  allow  the  sides  to  project  about  r-6-  of  an  inch.  Here 
we  have  lines  which  completely  represent  the  cavity  of  the  valve  and  the  valve  face. 
If  we  now  add  to  these  lines  the  proper  thickness  of  metal,  as  shown  in  Fig.  62,  this 
section  of  the  valve  will  be  complete. 

Fig.  63  shows  a  section  of  the  valve  taken  at  right  angles  to  that  shown  in  Fig.  62. 
Since  the  ports  are  14  inches  long,  the  cavity  of  the  valve  must  be  14  inches  wide,  as 


59 


shown.  Tlie  amount  that  the  valve  overlaps  the  ends  of  the  steam  ports  must  be  suffi- 
cient to  prevent  leakage.  For  a  valve  of  the  size  here  shown,  1  inch  overlap  is  allowed, 
and  tlic  thickness  of  metal  around  the  cavity  is  generally  £  of  an  inch.  For  smaller 
valves  the  overlap  at  each  end  of  the  steam  port  is  from  $  to  5  of  an  inch,  and  the 
thickness  of  metal  around  the  cavity  is  |  of  an  inch. 

The  valve  here  shown  is  suitable  for  a  locomotive  cylinder  16  inches  in  diameter, 
and  a  piston  speed  of  525  feet  per  minute,  and  the  dimensions  here  given  agree  with 
those  of  the  valves  that  are  at  present  in  use. 


INSIDE   LAP,    CLEARANCE   AND   INSIDE   LEAD. 


78.  Now,  a  few  words  in  regard  to  some  other  terms  used  in  connection  with  the 
slide-valve. 

INSIDE  LAP. — The  amount  that  the  inside  edges  i  and  lz  of  the  valve,  Fig.  65,  over- 
lap the  inside  edges  *  and  s.,  of  the  steam  ports,  when  the  valve  stands  midway  of  its 


Fig.  66 


travel,  is  called  inside  lap ;  thus,  the  distance  from  s  to  i,  or  from  s2  to  i.,,  represents  the 
inside  lap.  Its  purpose  is  to  delay  the  release  of  steam. 

The  amount  of  inside  lap  is  comparatively  small,  rarely  exceeding  &  of  an  inch, 
and  in  many  locomotives  the  valves  have  no  inside  lap.  Rules  for  determining 
the  inside  lap  cannot  be  given,  because  engineers  do  not  agree  on  this  subject.  The 
writer  believes  that  for  slow-running  locomotives,  particularly  if  these  have  to  run  over 
steep  grades,  a  little  inside  lap  will  be  beneficial.  For  ordinary  passenger  locomotives 
running  on  comparatively  level  roads,  no  inside  lap  should  be  used. 

7! i.  INSIDE  Ci.i :AKA\CE. — When  the  valve  stands  midway  of  its  travel,  as  shown  in 
Fig.  66,  and  its  inside  edges  i  and  /.,  do  not  cover  the  steam  ports,  then  the  amount  by 
which  each  edge  of  the  valve  comes  short  of  the  inner  edges  of  the  steam  ports  is  called 
inside  clearance;  thus,  the  distance  fromi  to  s,  or  from  12  to  s2>  represents  inside  clear- 
ance. The  purpose  of  inside  clearance  is  to  hasten  the  release,  and  is  sometimes 
adopted  in  very  fast-ninning  locomotives.  It  seldom  exceeds  /4  of  an  inch.  Good 
judgment  and  great  experience  are  required  for  determining  the  amount  of  clearance, 
and  in  deciding  for  what  classes  of  locomotives  it  should  be  used.  In  ordinary  pas- 
senger locomotives  the  valves  have  no  inside  clearance. 

SO.  The  widtli  of  opening  of  the  steam  port  for  the  release  of  steam  at  the  begin- 
ning of  the  stroke  is  called  inside  lead;  thus,  when  the  piston  is  at  the  beginning  of  its 
stroke,  and  the  valve  occupying  the  position  as  shown  in  Fig.  67,  then  the  distance 
between  the  inner  edges  /'.,  of  the  valve  and  the  inner  edge  s.,  of  the  steam  port  is  called 
inside  lead.  The  simple  terms  "load  "  and  "lap"  are  used  among  engineers  to  desig- 


60 


MODERN  LOCOMOTIVE  COXSTJIUCT1OX. 


nate  outside  lead  and  lap;  hence,  the  necessity  of  using  the  terms  "inside  lead"  and 
"  inside  lap  "  when  such  is  meant. 


THE   EVENTS   OP   THE  DISTRIBUTION   OF   STEAM. 

81.  In  the  distribution  of  steam  during  one  revolution  of  the  crank,  four  distinct 
events  occur,  namely : 

1st.  The  admission  of  steam. 

2d.  The  cutting  off,  or,  in  other  words,  the  suppression  of  steam. 

3d.  The  release  of  steam. 

4th.  The  compression  of  steam. 

Fig.  67.  The  outside  edges  c2  and  c3  of  the  valve,  and  the  outside  edges  o  and  o2  of 
the  steam  ports,  will  regulate  the  admission  and  suppression  of  steani ;  the  inner  edges 
i  and  i2  of  the  valve  and  the  inner  edges  s  and  s2  of  the  steam  ports  control  the  release 
and  compression  of  steam.  The  parts  of  the  stroke  of  the  piston  during  which  these 
events  will  happen  can  be  found  by  the  following  methods : 

EXAMPLE  25. — Travel  of  valve,  5  inches ;  lap,  1  inch ;  lead,  \  of  an  inch ;  stroke  of 
piston,  24  inches;  no  inside  lap  or  clearance.  Find  at  what  part  of  the  stroke  the 
admission,  suppression,  release,  and  compression  will  take  place. 

In  Figs.  67,  68,  and  69  the  valve  occupies  different  positions,  but  the  sections  of 
the  valve  in  these  figures  are  exactly  alike,  because  they  represent  one  and  the  same 


valve.  In  Fig.  67  the  distance  between  the  edge  c2  of  the  valve  and  the  edge  o  of  the 
steam  port  is  £  of  an  inch,  which  is  the  amount  of  lead  given  in  our  example ;  hence, 
this  position  of  the  valve  indicates  that  the  piston  is  at  the  beginning  of  its  stroke,  and 
the  angle  m  c  y  is  the  angular  advance  of  the  eccentric.  In  Fig.  68  the  edge  c.2  of  the 
valve  and  the  edge  o  of  the  steam  port  coincide,  and,  since  the  valve  is  moving  in  the 
direction  indicated  by  arrow  2,  the  suppression  commences,  or,  in  other  words,  the 
valve  is  cutting  off  steam  when  it  is  in  the  position  as  here  shown.  In  Fig.  69  the 
inside  edge  i  of  the  valve  coincides  with  the  inner  edge  s  of  the  steam  port,  and,  since 
the  valve  is  moving  in  the  direction  indicated  by  arrow  2,  the  release  must  commence 
when  the  valve  arrives  in  the  position  here  shown. 

In  Figs.  67,  68,  and  69  the  distances  from  the  outside  edge  o  of  the  steam  port  to 
the  center  c  of  the  circle  a  b  m  are  equal ;  that  is,  the  points  c  and  o  are  one  inch  apart, 
which  is  the  amount  of  lap.  The  diameters  of  the  circles  a  b  m  are  all  five  inches, 
which  is  the  travel  of  the  valve  given  in  the  example,  and  the  circumference  of  each 
circle  represents  the  path  of  the  eccentric,  and  also  the  path  of  the  center  of  the  crank- 
pin.  The  point  y  in  these  figures  represents  the  position  of  the  center  of  eccentric 
when  the  piston  is  at  the  beginning  of  its  stroke.  The  distance  between  the  point  y 


MOI>I:I;\  I.OI-OMOTI i '/•:  c(i\sn;rcTinN.  61 

.MI id  in  is  tli<>  same  in  all  figures,  and  consequently  the  angles  formed  by  the  lines  y  x 
and  in  <'  are  equal  and  represent  the  angular  advance  of  the  eccentric. 

When  the  valve  occupies  the  position  as  represented  in  Fig.  67,  the  center  line  of 
crank  will  coincide  with  the  line  A  Z?;  and  since  the  piston  will  then  be  at  the  beginning 
of  its  stroke,  it  follows  that  the  line  A  B  will  indicate  the  direction  in  which  the  piston 
must  move.  In  order  to  compare  the  relative  position  of  the  piston  with  that  of  the 
valve  with  as  little  labor  as  possible,  we  shall  assume  that  the  direction  in  which  the  pis- 
ton moves  is  represented  by  the  line  y  x,  instead  of  the  line  A  Z?;  hence  the  point  y  will 
not  only  show  the  position  of  the  center  of  the  eccentric,  but  it  will  also  indicate  the 
position  of  the  cent*  of  the  crank-pin  when  the  piston  is  at  the  commencement  of  its 
stroke.  If  these  remarks  are  thoroughly  understood,  there  will  be  no  difficulty  in  com- 
prehending that  which  is  to  follow. 

Now  let  us  trace  the  motions  of  the  valve  and  piston  and  thus  determine  at  what 
part  of  the  stroke  the  events  (previously  named)  will  take  place.  When  the  crank-pin 
is  moving  in  the  direction  as  indicated  by  the  arrow  marked  1,  Fig.  07,  the  center  of 
eccentric  will  move  through  part  of  the  circumference,  a  b  m,  and  the  valve  will  travel 
in  the  direction  indicated  by  the  arrow  2,  thus  opening  the  steam  port  wider  and 
wider  until  the  end  b  of  the  travel  is  reached ;  then  the  valve  will  commence  to  return, 
and  as  it  moves  toward  the  center  c,  the  steam  port  gradually  closes,  until  the  valve 
reaches  the  position  as  shown  in  Fig.  68 ;  then  the  steam  port  will  be  closed  and  steam 
cut  off.  To  find  the  position  of  the  piston  when  the  valve  is  cutting  off  steam,  we 
draw  through  the  edge  c.2  of  the  valve,  Fig.  68,  a  straight  line  c.,  y,  perpendicular  to  A 
B,  intersecting  the  circumference  a  b  m  in  the  point  //;  through  this  point  draw  a 
line  perpendicular  to  y  x  intersecting  the  latter  in  the  point  k,  and  this  point  k  being 
19J  inches  from  y  indicates  that  the  piston  has  traveled  19J  inches  from  the  beginning 
of  its  stroke  before  the  steam  is  cut  off,  and  that  steam  has  been  admitted  into  the 
cylinder  during  the  time  the  piston  traveled  from  y  to  k.  As  the  piston  continues  to 
move  towards  the  end  x  of  the  stroke,  the  valve  will  move  in  the  direction  of  the  arrow 
2,  Fig.  68,  and  the  steam  port  will  remain  closed  so  that  no  steam  can  enter  the  cylinder 
or  escape  from  it ;  hence  the  steam  that  is  now  confined  in  the  cylinder  must  push  the 
piston  ahead  by  its  expansive  force,  but  the 
moment  that  the  valve  reaches  the  position 
as  shown  in  Fig.  (i!)  the  release  of  steam  will 
commence.  To  find  the  corresponding  posi- 
tion  of  piston  we  draw  through  the  edge  c.,  of 
the  Valve,  Fig.  69,  a  line  c.,y,  perpendicular  to 
A  B,  intersecting  the  circumference  a  b  »i  in 
the  point  g.  Through  this  point  draw  a  line 
<l  I;  perpendicular  to  y  x,  intersecting  the  latter  in  the  point  k,  and  this  point  k 
being  22j|  inches  from  the  beginning  of  the  stroke  indicates  that  the  piston  has  trav- 
eled through  this  distance  when  the  release  of  steam  commences.  Now  notice,  the 
steam  is  cut  off  when  the  piston  has  traveled  19£  inches,  and  the  release  of  steam  com- 
mences when  the  piston  has  traveled  22;'  indies;  consequently  the  steam  is  worked 
expansively  during  the  time  the  piston  moves  ::;  indies  of  its  stroke.  The  steam  port 
will  remain  open  to  the  action  of  the  exhaust  during  the  time  the  piston  completes  its 


62  MODERN   LOCOMOTirE   CONSTRUCTION. 

stroke  and  moves  through  a  portion  of  its  return  stroke.  In  the  meantime  the  valve 
will  move  to  the  end  a  of  the  travel  and  return  as  indicated  by  arrow  4,  and  the  moment 
that  the  valve  again  reaches  the  position  shown  in  Fig.  69,  the  release  of  steam  will  be 
stopped.  To  find  the  corresponding  position  of  the  piston,  draw  through  the  edge  c2 
of  the  valve,  Fig.  69,  a  straight  line  c2  m  perpendicular  to  A  B,  intersecting  the  circum- 
ference a  b  m  in  point  m.  Through  this  point  draw  a  straight  line  m  k.,  perpendicular 
to  y  x,  and  intersecting  the  latter  in  the  point  k.,.  Since  the  distance  between  the 
points  x  and  k.z  is  22 1  inches,  it  follows  that  the  piston  has  moved  through  22f  inches 
of  its  return  stroke,  by  the  time  that  the  release  of  steam  will  cease.  As  the  valve 
continues  its  travel  in  the  direction  of  arrow  4,  Fig.  69,  the  ste^m  port  will  remain 
closed  until  the  edge  c2  of  the  valve  coincides  with  the  outer  edge  o  of  the  steam 
port,  and  during  this  time,  the  steam  which  remained  in  the  cylinder  is  compressed, 
but  as  soon  as  the  edge  c2  of  the  valve  passes  beyond  the  steam  port  edge  o,  the  admis- 
sion of  steam  into  the  cylinder  will  commence.  To  find  the  corresponding  position  of 
the  piston,  draw  through  the  outer  edge  o  of  the  steam  port,  Fig.  67,  a  straight  line 
o  g  perpendicular  to  A  B,  and  intersecting  the  circumference  a  b  m  in  the  point  g ; 
through  this  point  draw  a  line  g  k  perpendicular  to  y  x,  intersecting  the  latter  in  the 
point  k,  and  since  the  distance  between  the  points  x  and  k  is  23$  inches,  we  conclude 
that  the  piston  has  moved  through  23  5  inches  of  its  return  stroke  before  the  admission 
of  steam  will  begin.  Here  we  see  that  steam  will  be  admitted  into  the  cylinder  before 
the  return  stroke  of  the  piston  is  completed,  and  that  is  the  object  of  lead,  as  has  been 
stated  before.  Notice  once  more :  the  compression  of  steam  will  commence  when  the 
piston  has  traveled  22|  inches  of  its  return  stroke,  and  will  cease  when  the  piston  has 
traveled  23|  inches  of  its  return  stroke ;  hence  the  steam  is  compressed  during  the  time 
that  the  piston  travels  through  Ij  inches. 

In  each  one  of  these  figures  the  point  y  represents  the  relative  position  of  the 
center  of  eccentric  to  that  of  the  valve.  The  point  g  will  always  be  found  in  the  cir- 
cumference a  b  m,  and  in  a  straight  line  c2  g  drawn  perpendicular  to  A  B,  the  former 
passing  through  the  outer  edge  c2  of  the  valve. 

The  reason  why  the  point  y  should  in  all  cases  be  found  in  the  straight  line  c.2  g 
drawn  through  the  outside  edge  c2  of  the  valve  is  this :  the  center  c  of  the  circle  a  b  m 
has  been  placed  on  the  line  A  B  in  such  a  position  (as  shown  in  these  figures),  so  that 
the  distance  between  the  center  c  and  the  outside  edge  o  of  the  steam  port  is  equal 
to  the  lap,  therefore  the  center  g  of  the  eccentric  and  the  outer  edge  c2  of  the  valve 
will  always  lie  in  the  same  straight  line  drawn  perpendicular  to  A  B.  If  the  distance 
between  center  c  and  the  outer  edge  o  of  the  steam  port  is  greater  or  less  than  the 'lap, 
then  the  center  of  the  eccentric  and  outside  edge  of  the  valve  will  not  lie  in  the  same 
straight  line  drawn  perpendicular  to  the  line  A  B.  Here,  then,  we  can  conceive  the 
necessity  of  placing  the  center  c  of  the  circle  a  b  in  in  the  position  as  shown  in  these 
figures.  The  correctness  of  these  remarks  must  be  evident  to  the  reader  if  the  explana- 
tions in  the  previous  article  have  been  understood.  Again,  since  we  have  assumed  that 
the  point  g  not  only  represents  the  center  of  the  eccentric,  but  also  the  center  of  the 
crank-pin,  it  follows  that,  in  order  to  determine  how  far  the  piston  has  moved  from  the 
beginning  y  of  its  stroke  when  the  crank-pin  is  at  //,  we  must  draw  a  straight  lino 
through  the  point  <j  perpendicular  to  y  x,  as  has  been  done  in  these  figures.  See  Art.  69. 


MOnKK.\    LOCOMOTIVE   CONSTRUCTION. 


63 


From  those  constructions  we  can  obtain  our  answer  to  Example  25,  namely : 
Steam  will  be  cut  off,  or,  in  other  words,  suppression  will  commence  when  the 
piston  has  traveled  19i  inches  from  the  beginning  of  its  stroke,  and  steam  will  be 
admitted  into  the  cylinder  during  the  time  that  the  piston  travels  through  this  distance. 
The  steam  will  be  released  when  the  piston  has  traveled  22f  inches  from  the  beginning 
of  its  stroke,  consequently  the  steam  will  be  worked  expansively  during  the  time  the 
piston  travels  through  3£  inches.  The  release  of  steam  will  continue  until  the  com- 
pression commences,  which  will  occur  when  the  piston  has  traveled  22$  inches  of  its 
return  stroke.  The  compression  will  cease,  and  the  admission  of  steam  commence 
when  the  piston  has  traveled  23|  inches  of  its  return  stroke. 

The  same  answer  to  our  example  could  have  been  obtained  with  less  labor  by  a 
construction  as  shown  in  Fig.  70,  which  is  nothing  else  but  a  combination  of  the  three 
preceding  figures;  the  methods  of  finding  the  different  points  in  Fig.  70  have  not  been 
changed,  and  therefore  an  explanation  in  connection  with  this  figure  is  unnecessary. 


POWER    REQUIRED    TO    WORK    PLAIN    SLIDE-VALVE,    ROLLER    VALVES    AND    BALANCED    SLIDE- 
VALVES. 

82.  The  great  aim  of  engineers  in  constructing  a  slide-valve  for  a  locomotive  is  to 
produce  a  valve  that  will  require  as  little  power  as  possible  to  work  it.     Consequently, 


y  Bfyinntng  of 
" 


Fig.  70 


when  a  plain  slide-valve  is  to  be  used,  such  as  we  have  shown  in  some  of  the  foregoing 
articles,  the  valve  face  is  made  as  small  as  it  can  be  made,  without  interfering  with 
the  duties  which  the  valve  has  to  perform. 

Again,  it  is  of  very  great  importance  to  have  the  mechanism  inside  of  a  steam- 
chest  as  simple  as  possible;  therefore,  on  account  of  the  great  simplicity  of  the  plain 
slide-valve,  it  is  more  extensively  used  than  any  other  kind.  Yet  it  has  its  drawbacks, 
besides  requiring  a  considerable  amount  of  .force  to  move  it  forward  and  backward  on 
its  seat;  for  instance,  it  is  liable  to  cut  the  valve  sent,  wear  the  link,  rocker,  pins,  and 
eccentrics  comparatively  fast,  and  when  the  valve  is  large,  it  is  difficult  to  reverse  the 
engine. 

Let  us  for  a  moment  consider  the  amount  of  force  that  will  be  required  to  move 
an  ordinary  plain  slide-valve.  The  resistance  which  must  be  overcome  in  moving  any 


(54  MODERN  LOCOMOTIVE    CONSTRUCTION. 

slide-valve  is  simply  the  friction  between  the  valve  and  its  seat.  This  friction  depends 
upon  the  pressure  of  the  valve  against  the  seat,*  and  this  pressure  is  equal  to  the  total 
steam  pressure  iipon  the  back  of  the  valve,  minus  the  reaction  of  the  steam  pressure 
in  the  steam  and  exhaust  ports. 

This  state  of  affairs  we  have  endeavored  to  illustrate  in  Fig.  71,  in  which  the 
arrows  marked  1  indicate  the  pressure  of  the  steam  on  the  back  of  the  valve,  and  the 
arrows  marked  2  indicate  the  reaction  of  the  steam  pressure  in  the  ports. 

To  present  in  a  plain  manner  the  subject  of  finding  the  amount  of  this  friction, 
let  us  take  the  following  example:  The  valve  is  9  inches  long  (see  Fig.  71)  and  16 
inches  wide ;  the  steam  pressure  in  steam-chest  is  120  pounds,  it  is  required  to  find  the 
amount  of  force  necessaiy  to  move  the  valve. 

The  total  pressure  on  the  back  of  the  valve  is  found  by  multiplying  the  area  of  the 
valve  face  by  the  steam  pressure,  hence  we  have  9  x  16  =  144  inches,  which  is  the  area 
of  the  valve  face,  and  144  x  120  =  17,280  pounds,  which  is  the  total  pressure  on  the 
back  of  the  valve,  but  it  is  not  the  pressure  of  the  valve  against  its  seat. 

To  obtain  the  latter,  we  must  deduct  from  the  total  pressure  on  the  back  of  the 
valve  the  reacting  pressure  in  the  ports.  The  reaction  of  the  steam  pressure  in  the 
ports  can  only  be  obtained  approximately,  because  there  are  no  data  from  which  we 
can  make  the  calculation.  It  will  readily  be  seen  that  the  reaction  of  the  steam  press- 
ure in  the  ports,  when  the  valve  is  in  full  gear,  is  affected  by,  and  depends  upon,  the 
size  of  the  exhaust  nozzle,  the  speed  of  engine,  the  lap  of  valve,  and  some  other  details. 

This  pressure  is  also  variable  during  the  travel  of  the  valve.  For  our  purpose  here, 
we  will  take  the  average.  From  observation  the  writer  is  led  to  believe  that  in  ordi- 
nary locomotives  the  reacting  pressure  is  equal  to  one-third  of  the  pressure  on  the  back 
of  the  valve.  Grant  that  this  is  true,  then  the  total  pressure  acting  in  the  direction  of 
the  arrows  2  will  be  equal  to  J^-f  —  =  5,760  pounds.  Subtracting  this  quotient  from  the 
total  steam  pressure  on  the  back  of  the  valve,  we  obtain  the  pressure  of  the  valve 
against  its  seat,  hence  we  have  17,280  —  5,760  =  11,520  pounds,  which  is  the  pressure  of 
the  valve  against  the  seat. 

In  order  that  the  reader  may  obtain  a  better  idea  of  the  effect  of  this  pressure,  we 
may  say  that  this  valve  has  to  be  moved  forward  and  backward  with  a  load  of  11,520 
pounds  (nearly  six  tons)  upon  its  back. 

The  friction  between  two  cast-iron  surfaces  which  are  straight,  smooth,  and  lubri- 
cated generally  ranges  from  n,  to  yj-  of  the  pressure.  In  this  case  we  will  adopt  the 
former  proportion.  Therefore,  the  friction  between  the  valve  and  seat  is  found  by 
dividing  the  pressure  on  the  valve  seat  by  10,  hence  •uflf^L  =  1,152  pounds. 

This  1,152  pounds  is  the  friction,  or,  we  may  say,  the  resistance  which  must  be 
overcome  in  moving  the  valve,  therefore  a  force  of  1,152  pounds  is  required  to  move 
the  valve  on  its  seat  in  a  direction  as  indicated  by  arrow  3,  and  the  same  amount  of 
force  is  also  required  to  move  the  valve  in  azi  opposite  direction.  To  work  such  a  valve 
as  quickly  as  must  be  done  in  a  locomotive  will  need  a  great  amount  of  power ;  also, 
whatever  power  is  used  for  this  purpose  is  a  loss,  because  the  engine  will  have  that 
amount  less  with  which  to  perform  useful  work,  that  is,  to  haul  the  train. 

It  nmst  also  be  plain  to  the  reader  that  a  valve  working  under  such  a  pressure  is 

*  The  weight  of  the  valve  being  comparatively  small,  it  is  left  out  of  the  question. 


M01>Ki:\ 


COXSTSrCTIOA'. 


65 


very  liable  to  cut  the  valve  seat,  and  the  only  way  that  this  evil  may  be  prevented  to 
some  extent  is  to  use  the  best  of  metal  in  the  cylinders,  and  keep  the  valve  well  oiled. 
We  have  shown  that  this  valve  requires  a  considerable  amount  of  force  to  move  it  on 
its  seat,  and  since  this  force  is  transmitted  through  the  eccentrics,  links,  rockers,  and 
l>ii is,  it  follows  that  these  will  also  wear  very  quickly.  Again,  to  pull  the  reverse  lever 
from  one  notch  to  another,  on  a  locomotive  having  cylinders  about  16  inches  in  diame- 
ter, is  often  laborious  work  for  an  engineer,  and  still  more  so  in  modern  engines, 
because  these  generally  have  larger  cylinders,  and  consequently  larger  valves. 

From  the  foregoing  remarks  we  infer  that  although  the  plain  slide-valve  is 
extremely  simple,  and  can  and  must  be  made  to  distribute  the  steam  correctly,  the 
steam  pressure  on  the  back  of  this  valve  impairs  its  usefulness;  and  consequently  there 
is  a  desire  existing  among  engineers  to  procure  a  valve  in  which  the  evil  effects  of 
pressure  on  the  back  of  the  same  will  be  removed. 


ROLLER  VALVES. — BALANCED  VALVES. 

83.  There  are  in  use  two  distinct  kinds  of  valves  which  require  less  power  to  work 
than  the  plain  slide-valve.  One  kind  comprises  the  roller  valves,  and  to  the  other  kind 
belong  the  balanced  valves,  or,  as  sometimes  called,  the  equilibrium  slide-valve. 

Although  less  power  is  required  to  work  the  roller  valve  than  is  needed  for  the 
plain  slide-valve — and  therefore  when  a  roller  valve  is  used  the  wear  of  the  valve  gear 
will  be  reduced — this  valve  has  never  been,  and  is  not  now  veiy  extensively  used,  and, 
in  the  writer's  opinion,  this  should  not  be  a  matter  of  surprise,  because  in  the  con- 
struction of  these 


Fiff.   72  |      Steam  chest  cover 


'^•iFtff.73 


Fig. 


valves,  no  at- 
tempt has  been 
made  to  remove 
the  steam  press- 
ure on  the  back  of 
the  valve,  which 

so  impaired  the  usefulness  of  the  plain  slide- 
valve.  But  since  the  roller  valve  is  sometimes 
adopted,  the  writer  believes  that  a  description 
of  it  will  be  interesting  to  the  reader. 

Figs.  72,  73,  and  74  represent  different  views 
of  this  valve,  and,  as  will  be  seen,  the  only  dif- 
ference between  the  roller  valve  and  the  plain 
slide-valve,  is  that  the  former  is  constructed 
in  such  a  manner  as  to  make  room  for  the  rollers  r  r  r,  which  are  interposed  between 
the  steam-chest  seat  and  the  valve.  These  rollers  are  prevented  from  touching  each 
other  by  the  small  axles  x  x  x,  and  around  these  the  rollers  turn.  These  axles,  with 
the  rollers  placed  upon  them,  are  riveted  to  the  bars  t  t,  so  that  when  this  arrange- 
ment is  completed,  as  shown  in  Fig.  74,  a  small  carriage  is  obtained  on  which 
the  valve  is  to  work.  To  prevent  wear  of  the  steam-chest  seat  and  valve  as  much 
as  possible,  steel  plates  p  p,  Figs.  72  and  73,  are  laid  on  the  steam-chest  seat,  but 


66 


MODERN  LOCOMOTirfi   CONSTRUCTION. 


not  fastened  to  it.  Two  other  plates  u  u  are  attached  to  the  valve,  and  between  these 
plates  the  carriages  are  made  to  roll.  The  most  important  dimensions  are  given  in  the 
figures,  and  the  general  arrangement  is  also  plainly  shown,  so  that  a  further  description 
is  unnecessary. 

84.  A  balanced  slide-valve  is  represented  in  Figs.  75,  76,  and  77.     This  valve  is 
extensively  used  in  modern  locomotives,  and  seems  to  be  growing  in  favor.     In  the 
construction  of  these  valves  the  correct  principle  has  been  followed,  namely,  the 
removal  of  the  steam  pressure  on  the  back  of  the  valve.     This  is  accomplished  by  cut- 
ting grooves  around  the  back  of  the  valve,  and  care  must  be  taken  to  cut  these  grooves 
perfectly  true ;  in  these,  strips  s  s  s  of  cast-iron  are  accurately  fitted,  so  that  the  latter 
may  move  up  and  down  in  the  grooves  without  any  perceptible  play.     The  strips  are 
held  up  by  spiral  springs  t  1 1,  as  shown.     Some  mechanics  make  these  springs  of  hard 
brass  wire,  while  others  prefer  to  use  steel  wire. 

The  diameter  of  the  wire  ranges  from  -^  to  £  of  an  inch.  The  number  of  springs 
generally  employed  is  shown  in  the  figures.  When  the  balanced  valve  is  to  be 
used  the  under  side  of  the  steam-chest  cover  must  be  accurately  planed  and  scraped, 
because,  against  this  surface,  the  strips  must  press  and  make  a  steam-tight  joint  when 
the  valve  is  in  the  steam-chest,  as  shown  in  Fig.  78.  This  arrangement  will  prevent 
the  steam  from  coming  into  contact  with  the  greater  portion  of  the  back  of  the  valve, 

thus       reducing 
L ^_^_^ ,g  nwniKTwt^j  ---  ---.*,„ ».w..,," v.~.      the   pressure   of 

the  valve  against 
the  valve  seat, 
and  therefore 
less  power  will 
be  required  to 
work  it  than 
would  be  needed 

for  a  plain  slide-valve,  and  the  wear  in  the 
valve  gear  will  also  be  reduced. 

When  the  valve  is  new,  and  placed  in 
the  steam-chest,  as  shown  in  Fig.  78,  and 
all  ready  for  use,  then  the  strips  should 
project  not  more  than  -^  of  an  inch  above 
the  top  of  the  valve.  The  strips  should 

Fia.  77  \D      Section  through  A'.*:  AVI  xi  -IQ-T-         j  r>  •       1, 

not  be  less  than  If  inches  deep ;  2  inches 

is  better;  in  fact,  these  should  be  made  as  deep  as  they  can  be  made  without 
choking  the  exhaust  cavity  of  the  valve;  and  %  inch  is  about  the  proper  width  of 
the  strips.  The  manner  of  joining  them  at  the  corners  is  plainly  shown  in  the  plan 
of  the  valve,  Fig.  77.  For  the  sake  of  distinctness  the  strips  have  been  shaded  in  this 
figure. 

85.  Some  master  mechanics  object  to  the  spiral  springs,  because  the  pockets  in 
which  these  springs  are  placed  will  in  time  fill  up  with  tallow,  and  thus  prevent  the 
springs  from  working  freely.     Therefore,  in  place  of  spiral  springs  for  holding  the 
strips,  elliptic  springs  made  of  flat  steel  are  used,  as  shown  in  Figs.  79  and  80,  and,  of 


VO/I/.7.-.Y 


ro.v.x  run  •  n<>\. 


67 


course,  when  these  springs  are  to  be  used,  the  grooves  must  be  cut  deep  enough  to 
receive  them,  and  the  pockets,  as  shown  in  Figs.  75  and  76,  left  out. 

Yet  these  elliptic  springs  are  not  free  from  objections.  Some  master  mechanics 
disapprove  of  them  because  they  are  too  strong  when  new,  and  consequently  are  liable 
to  cut  tlic  steam-chest  cover. 

On  the  other  hand,  if  these  elliptic  springs  are  made  weaker,  so  that  they  will  work 
satisfactorily  in  the  beginning,  they  will  not  remain  so  long,  and  become  too  weak  to 


Fig.  78 

hold  up  the  strips.  The  result  of  this  disagreement,  among  master  mechanics,  is  that 
there  are  about  as  many  valves  wth  spiral  springs  in  use  as  there  are  with  elliptical 
springs. 

All  balanced  valves  of  this  kind  should  have  a  hole  /«,  about  f  of  an  inch  in  diame- 
ter, through  the  top  of  the  valve;  without  this  hole  the  valve  will  lift  off  its  seat. 
When  a  locomotive  is  running  and  then  steam  shut  off,  a  partial  vacuum  will  be  formed 
in  the  steam-chest,  causing  the  valve  to  chatter,  and  thus  ruin  its  mechanism.  To  pre- 
vent the  forming  of  a  vacuum  in  the  steam-chest,  and  consequently  its  evil  effects,  a 
vacuum  valve  must  be  attached  to  the  steam-chest  to  admit  air  into  the  latter  when 
steam  is  shut  off  during  the  time  the  engine  ts  in  motion.  Without  this  vacuum  valve 


Fly.  ?!t 


Fig.  SO 


no  balanced  slide-valve  must  be  expected  to  work  successfully.  Indeed,  it  is  good 
practice  to  attach  a  vacuum  valve  to  all  steam-chests  in  which  a  plain  slide-valve  is 
working,  as  this  will  often  be  the  means  of  preventing  the  ashes  from  being  drawn 
into  the  cylinder  and  the  steam-chest. 

When  these  balanced  valves  are  properly  made  by  mechanics,  not  cheap  labor,  good 
results  can  certainly  be  looked  for.  We  find  it  recorded  in  the  annual  report  of  the 
"American  Railway  Master  Mechanics'  Association,"  for  the  year  1884,  that  a  passenger 
locomotive  with  16"  x  22"  cylinders,  driving  wheels  .">4  feet  in  diameter,  fitted  with  Morse 
balanced  valves,  ran  166,000  miles  without  any  need  of  facing  the  valve  seat. 

Another  good  result  obtained  by  using  the  balanced  valves,  is  that  the  reverse 
level-  can  easily  be  handled,  something  to  be  appreciated  by  the  engineer  having  charge 
of  the  engine.  The  balanced  valve,  which  of  late  seems  to  be  the  favorite,  is  shown 
in  the  figures,  and  is  called  the  "Richardson"  balanced  slide-valve.  But  by  this 
remark  the  writer  does  not  wisli  to  be  understood  that  this  is  the  only  good  balanced 
valve;  he  simply  wishes  to  state  facts  as  they  appeal-  to  him. 


68 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


ALLEN  VALVE. 

86.  The  Allen  valve  is  shown  in  Fig.  BOA.  Its  general  design  is  the  same  as  that 
of  an  ordinary  D  valve  with  the  exception  that  it  has  a  supplementary  steam  passage 
P  P  cast  into  it. 

In  this  valve,  as  in  the  ordinary  D  valve,  the  outside  lap,  or  simply  lap,  is  equal  to 
a  g,  that  is  to  say,  it  is  equal  to  the  distance  by  which  it  overlaps  the  outer  edge  y  of 

the  steam  port,  and  this  lap  is  in  no- 
wise affected  by  the  supplementary 
steam  passage ;  in  fact,  all  the  defini- 
tions of  lap,  lead,  linear  advance  of 
valve,  angular  advance  of  eccentric, 
etc.,  remarks  and  rules  given  in  con- 
nection with  the  plain  slide-  valve,  are 
not  changed  when  the  Allen  valve  is 
used. 

Fig.  80c  shows  the  valve  at  the 
end  of  its  travel.  It  will  here  be  no- 
ticed that  the  thickness  a  f  of  the 
metal  outside  of  the  supplementary 
port  covers  a  portion  of  the  steam 
port  S,  and  therefore  a  somewhat  wider  steam  port  may  be  required  than  for  an  ordi- 
nary slide-valve. 

To  find  the  proper  width  of  steam  port  for  an  Allen  valve  under  these  conditions, 
we  should  proceed  as  follows : 
First,  find  the  width  of  steam 
port  required  for  a  free  ex- 
haust for  an  ordinary  slide- 
valve,  as  explained  in  Art.  43. 
Now  referring  to  Art.  59,  we 
find  that  for  the  admission  of 
steam  we  require  only  n,  of 
this  width  of  port  ;  conse- 
quently, if  the  thickness  of  the 
wall  a  /is  greater  than  -f-0  of 
the  width  of  the  port,  we  must  make  the  latter 
correspondingly  wider.  To  illustrate: 

If  we  find  by  computation,  as  explained  in 
Art.  43,  that  the  width  of  the  steam  port  for  a  free 
exhaust  should  be  1J  =  1.25  inches,  then  for  the  ad- 
mission of  steam  we  shall  require  an  opening  of 
1.25  x  .8  =  1  inch.  If,  now,  the  thickness  of  the  wall 
a /is  f  inch,  then  we  shall  require  for  the  Allen 
valve  a  steam  port  1  +  g  =  1|  inches  wide.  This  shows  us  that  in  this  particular  case  the 
width  of  steam  port  for  an  Ah1  en  valve  should  be  £  inch  greater  than  for  a  plain  valve, 


4L_  '* 

K 

i    , 

" 

-M 

in 

i  .  . 

J 

'Iff 

™T 

I 

[J  f-  ill    .. 

—  i 

U 

r~    s 
/      7 

—  J. 

• 

18>f-|- 


Fig.  80  D. 


MODERX  LOCOMOTIVE   COXSTRUCTI<>\.  QQ 

When  the  correct  width  of  steam  port  has  been  found  as  above,  and  the  inner  edge  c, 
of  tho  supplementary  port  is  to  be  in  line  with  the  inner  edge  of  the  steam  port  when 
the  valve  is  at  the  end  of  its  travel,  as  shown  in  Fig.  80c,  then  the  travel  of  the  valve 
will  lie  c(|iial  to  twice  the  sum  of  the  width  of  the  steam  port  plus  the  amount  that  the 
edge  r  overlaps  the  outer  edge  of  port.  It  will  be  noticed  that  this  rule  differs  slightly 
from  that  given  in  Art.  59  for  finding  the  travel  of  an  ordinary  slide-valve.  The  cor- 
rertness  of  this  travel  should  be  checked  by  drawing  the  valve  seat  with  the  valve  at 
the  cud  of  its  travel,  as  shown  in  Fig.  80c.  If  in  this  position  the  steam  port  is  fully 
or  very  nearly  uncovered  so  as  to  give  a  free  exhaust,  the  travel  as  previously  found  is 
correct.  (lencrally,  when  the  steam  ports  are  correctly  proportioned,  the  travel  of 
an  Allen  valve  will  bo  less  than  that  of  an  ordinary  valve. 

The  arrangement  of  the  valve  and  seat  should  be  such  that  when  the  outer  edge  a 
(Fig.  S()B)  of  the  valve  admits  steam  into  the  cylinder,  then  the  edge  b  of  the  supple- 
mentary port  should  also  admit  steam  into  the  same  end  of  cylinder;  the  flow  of  steam 
under  these  conditions  is  indicated  by  the  arrows.  To  attain  this  object,  it  is  necessary 
to  assign  to  the  valve  seat  a  correct  length,  which  is  done  by  making  a  drawing  of  the 
seat  and  the  valve  in  its  central  position,  as  shown  in  Fig.  80x,  and  then  making  b  h 
equal  to  a  ff,  the  lap. 

The  following  advantage  is  claimed  for  this  valve :  In  high  speed  locomotives  with 
the  link  well  hooked  up,  say,  so  as  to  cut  off  at  6  inches  of  the  stroke,  the  greatest 
width  of  steam  port  opening  with  an  ordinary  valve  is  about  jj  of  an  inch  only; 
with  this  contraction  great  difficulty  is  often  experienced  to  keep  up  a  full  steam 
pressure  from  the  beginning  of  the  stroke  of  piston  to  the  point  of  cut-off ;  in  fact, 
diagrams  taken  under  these  conditions  always  show  a  marked  fall  of  steam  pressure 
during  this  period. 

If,  now,  an  Allen  valve  is  used  with  supplementary  ports  £  inch  wide,  then,  instead 
of  having  a  steam  port  opening  of  |  inch  in  width,  as  will  be  the  case  when  the  ordi- 
nary valve  is  used,  we  shall  have  a  port  opening  of  double  that  amount,  which  gives  a 
freer  admission  of  steam,  and  consequently  the  engine,  with  its  links  hooked  up,  will 
be  capable  of  making  better  time,  and  in  some  cases  do  the  work  with  less  fuel.  But 
the  fact  that  this  valve  is  not  universally  adopted  seems  to  indicate  a  want  of  confi- 
dence in  its  advantages ;  indeed,  we  have  heard  mechanics  of  ability  express  the  opinion 
that  the  two  currents  of  steam  flowing  into  the  same  steam  port  will  interfere  with 
each  other's  flow,  thereby  losing  the  advantage  gained  by  an  increased  port  opening. 

Fig.  80D  shows  different  views  and  details  of  a  balanced  Allen  valve.  It  is  used  in 
an  18  x  24  inch  passenger  engine. 

TO  FIND  THE  POINT  OF  CUT-OFF  WHEN   LENGTH  OF  CONNECTING-ROD  IS  GIVEN. 

87.  During  one  revolution  of  the  crank  the  piston  makes  two  strokes;  one  stroke 
we  will  call  the  forward  stroke,  and  the  other,  the  return  stroke.  When  the  length  of 
the  connecting-rod  is  assumed  to  be  infinite,  and  the  valve  constructed  according  to  the 
foregoing  methods,  linving  the  same  amount  of  lap  at  each  end,  then  the  valve  will  cut 
off  equal  portions  of  steam  in  the  forward  and  return  strokes. 

When  a  connecting-rod  of  definite  length  is  introduced,  instead  of  a  rod  whose 


70 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


length  is  assumed  to  be  infinite,  but  leaving  everything  else  unchanged,  then  the  por- 
tions of  steam  cut  off  in  the  forward  stroke  will  not  be  equal  to  that  cut  off  in  the 
return  stroke.  To  illustrate  this,  let  us  take  the  following  example : 

EXAMPLE  26. — The  length  of  the  connecting-rod  is  4  feet  (this  short  length  has  been 
adopted  for  the  sake  of  clearness) ;  stroke  of  piston  is  24  inches ;  lap  of  valve,  l£  inches ; 
no  lead,  and  travel  of  the  valve  5J  inches ;  find  at  what  part  of  the  forward  stroke,  and 
also  at  what  part  of  the  return  stroke,  steam  will  be  cut  off. 

In  order  not  to  complicate  matters,  we  will  assume  the  length  of  the  eccentric-rod 
to  be  infinite.  This  problem  will  be  divided  into  two  parts.  1st.  It  will  be  shown  at 

what  part  of  the 
forward,  and  also 
at  what  part  of 
the  return  stroke, 
steam  will  be 
cut  off  when  the 
length  of  the  con- 
necting-rod is  as- 
sumed to  be  in- 
finite. 

2d.  It  will  be 
shown     at    what 

part  of  the  forward,  and  also  at  what  part  of  the  return 
stroke,  steam  will  be  cut  off  when  the  connecting-rod  is 
4  feet  long. 

Fig.  81.  Draw  a  straight  line,  A  B ;  on  this  lay  off  the 
exhaust  and  steam  ports ;  draw  the  section  of  the  valve  so 
that  it  will  overlap  each  steam  port  l£  inches,  as  shown ; 
drawing  the  valve  in  this  position,  we  represent  it  to  be  in 
its  central  position.  Take  the  edge  c  of  the  valve  as  a 
center  and  with  a  radius  of  2£  inches  (equal  to  half  the 
travel)  draw  the  circle  a  b  m.  The  circumference  of  this 
circle  will  represent  the  path  of  the  center  of  the  eccentric, 
and  also  that  of  the  crank-pin  (see  Example  19). 

The  direction  in  which  the  crank-pin  is  to  move  is  indicated  by  the  arrow  marked  1. 
Since  there  is  to  be  no  lead,  draw  through  the  outside  edge  o  of  the  steam  port  a 
straight  line  i  h  perpendicular  to  A  J5,  intersecting  the  circumference  a  b  m  in  the  point 
y  and  y.  The  point  y  will  be  the  center  of  the  eccentric,  and  also  the  center  of  the 
crank-pin  when  the  piston  is  at  the  beginning  of  its  forward  stroke,  and  the  point  g 
will  represent  the  center  of  the  eccentric  and  that  of  the  crank-pin  at  the  moment  that 
steam  is  cut  off  in  the  forward  stroke.  Through  the  points  y  and  c  draw  the  diameter 
y  x,  which  will  represent  the  stroke  of  the  piston ;  divide  y  x  into  24  equal  parts ;  each 
part  will  then  represent  one  inch  of  the  piston's  stroke.  The  direction  in  which  the 
piston  moves  during  the  forward  stroke  is  indicated  by  the  arrow  2,  and  consequently, 
in  the  return  stroke  the  piston  must  move  in  the  direction  as  indicated  by  the  arrow  3. 
Since  y  is  the  beginning  of  the  forward  stroke,  we  commence  at  y  in  marking  the 


M»in-:i;.\  i.ucoMornE  m.\sn;i  <  TION.  71 

inches  '2,  4,  6,  etc.,  on  the  right-hand  side  of  x  y;  and  because  x  is  the  beginning  of 
tlic  return  stroke,  we  commence  at  ./•  in  marking  the  inches  2,  4,  6,  etc.,  on  the  left-hand 
side  of ./'  /;. 

Through  the  point  y  draw  a  straight  line  perpendicular  to  y  x,  intersecting  the 
latter  in  the  point  /.-,  which  is  found  to  be  located  18f  inches  from  the  point  y,  therefore 
tin-  piston  will  travel  18jf  inches  from  the  beginning  of  its  forward  stroke  before  steam 
is  cut  off.  So  far,  the  construction  is  similar  to  that  shown  in  Fig.  5o,  and  explained 
in  Kxample  l!>.  We  have  assumed  that  the  diameter  //  ./•  represents  the  stroke  of  the 
piston,  the  point  //  the  beginning  of  the  forward  stroke,  aud  the  point  x  the  beginning 
of  the  return  stroke.  But  in  these  constructions,  the  center  of  the  crank-pin  and  the 
center  of  eccentric  are  always  assumed  to  be  represented  by  the  same  point,  therefore 
the  point  x  will  also  represent  the  position  of  the  center  of  eccentric,  when  the  piston 
is  at  the  beginning  of  the  return  stroke.  Since  the  amount  of  lap  is  the  same  at  either 
end  of  the  valve,  the  center  of  eccentric  must  travel  from  the  beginning  x  of  the  return 
stroke  through  an  arc  equal  to  the  arc  y  <j  in  order  to  reach  the  position  at  which  the 
steam  will  be  cut  off  in  this  stroke ;  therefore  the  arc  x  fj2  must  be  made  equal  to  the 
arc  y  y.  The  best  way  to  accomplish  this  is  to  draw  a  line  i,  h2  through  #,  and  per- 
pendicular to  A  B,  intersecting  the  circumference  a  b  m  in  the  point  ff2,  then  the  arc 
x  y.,  will  be  equal  to  the  arc  y  y,  and  the  point  y.2  will  be  the  position  of  the  center  of 
eccentric  at  which  steam  will  be  cut  off  in  the  return  stroke. 

Through  the  point  y.,  draw  a  straight  line  y2  r  perpendicular  to  y  x,  cutting  the 
latter  in  the  point  r.  The  distance  from  x  to  r  will  represent  the  portion  of  the  stroke 
during  which  steam  will  be  admitted  into  the  cylinder,  and  since,  according  to  our  con- 
st nu-tion,  the  point  r  is  situated  18f  inches  from  x,  it  follows  that  the  piston  will  reach 
the  point  of  cut-off  when  it  has  traveled  18§  inches  from  the  beginning  of  the  return 
stroke.  Here,  then,  we  see  that  when  the  length  of  the  connecting-rod  is  assumed  to 
be  infinite,  steam  will  be  admitted  into  the  cylinder  during  equal  portions  of  the  two 
strokes,  or,  in  other  words,  the  distance  from  the  beginning  of  the  forward  stroke  to 
the  point  of  cut-off  will  be  equal  to  that  in  the  return  stroke.  Now  let  us  consider 
at  what  part  of  the  stroke  steam  will  be  cut  off  when  the  connecting-rod  is  4  feet 
long. 

If  the  valve  is  drawn  full  size  in  the  construction,  then  the  diameter  y  x  will  be  5J 
inches  long;  but  we  have  assumed  that  this  diameter  also  represents  the  length  of  the 
stroke  of  piston,  which  is  2  feet;  therefore,  when  we  lay  off  the  length  of  the  connect- 
ing-rod, we  must  adopt  the  diameter  y  x  as  a  scale  2  feet  long,  consequently  one-half 
of  this  diameter  (which  is  equal  to  the  radius  of  the  circle  a  b  m)  will  represent  one 
foot,  and  the  length  of  the  connecting-rod,  which  is  4  feet  long,  will  be  equal  to  four 
times  the  radius  of  the  circle  n  b  ///.  Prolong  the  line  //  x  to  any  length,  say,  to  7),  then 
the  path  of  the  cross-head  pin  will  lie  in  the  line  y  I),  or,  in  other  words,  the  center  of 
cross-head  pin  will  always  be  found  somewhere  in  the  line  y  D.  From  the  point  y  as  a 
center,  and  with  a  radius  ei|iial  to  the  length  of  the  connecting-rod  (equal  to  four  times 
the  radius  of  the  circle  a  //  ///).  describe  an  arc  cutting  the  line  //  />  in  the  point  //.,;  this 
point  will  represent  the  position  of  the  center  of  the  cross-head  pin  when  the  piston  is 
at  the  beginning  of  its  forward  stroke.  From  the  point  ./  as  a  center,  and  with  a  radius 
equal  to  the  length  of  the  connecting-rod,  describe  an  arc  cutting  the  line  //  I)  in  the 


72 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


point  #2,  and  this  point  will  represent  the  center  of  the  cross-head  pin  when  the  piston 
is  at  the  end  of  the  forward  stroke  or  the  beginning  of  the  return  stroke. 

The  distance  between  the  points  yz  and  x2  will  represent  the  length  of  stroke,  and  will 
be  equal  to  the  diameter  y  x.  Divide  the  distance  between  y2  and  x2  into  twenty-four 
equal  parts  and  number  them  as  shown.  The  arrow  marked  4  indicates  the  direction 
in  which  the  piston  moves  during  the  forward  stroke,  and  arrow  5  indicates  the  direc- 
tion of  the  motion  of  the  piston  during  the  return  stroke.  From  the  point  g  in  the 
circumference  a  I  m  as  a  center,  and  with  a  radius  equal  to  the  length  of  the  con- 
necting-rod, describe  an  arc  cutting  the  line  y  D  in  the  point  p ;  this  point  p  will 
be  the  center  of  the  cross-head  pin,  when  steam  is  cut  off  in  the  forward  stroke  of  the 
piston,  and,  as  will  be  seen  in  the  figure,  the  point  p  is  situated  midway  between  the 
divisions  marked  17  and  18,  and  therefore  indicates  that  steam  will  be  cut  off  when  the 
piston  has  traveled  17J  inches  from  the  beginning  of  the  forward  stroke  yz. 

From  the  point  fj.2  (in  the  circumference  a  b  tn)  as  a  center,  and  with  a  radius  equal 
to  the  length  of  the  connecting-rod,  describe  an  arc  cutting  the  line  y  D  in  the  point  s ; 
this  point  s  will  be  the  position  of  the  center  of  the  cross-head  pin  when  steam  is  cut  off 
in  the  return  stroke,  and,  as  will  be  seen  in  the  figure,  it  is  situated  19f  inches  from  the 
beginning  of  the  return  stroke  x2,  and  therefore  indicates  that  steam  will  be  cut  off  in 
the  return  stroke  when  the  piston  has  traveled  19f  inches  from  x2.  Here,  then,  we  see 

that  when  the  connecting-rod  is  4  feet  long,  steam  will  be  cut  off 
in  the  forward  stroke  when  the  piston  has  traveled  17£  inches, 
and  in  the  return  stroke  steam  will  be  cut  off  when  the  piston 
has  traveled  19f  inches,  making  a  difference  of  2  J  inches  between 
the  two  points  of  cut-off.  This  difference  is  caused  by  the  an- 
gularity of  the  connecting-rod,  or,  in  other  words,  by  the  angle 
formed  between  the  center  line  of  the  connecting-rod  and  the 
line  y  D.  This  angle  can  be  reduced  by  making  the  connecting- 
rod  longer,  but  not  changing  the  length  of  the  stroke ;  with  this 
change  the  difference  in  position  between  the  points  of  cut-off 
in  the  forward  and  return  strokes  will  be-  decreased.  But  in  all 
engines  in  which  the  valve  receives  its  motion  direct  from  the 
eccentric,  with  an  equal  amount  of  lap  and  lead  at  each  end 
of  the  valve,  there  will  always  be  a  difference  in  the  position  of 
the  points  of  cut-off,  even  if  the  connecting-rod  is  comparatively 
long. 

Should  it  be  desirable  to  make  the  valve  in  these  engines  to 
cut  off  equal  portions  of  steam  in  the  return  and  forward  strokes,  then  the  only  way 
that  this  can  be  accomplished  is  by  giving  the  valve  more  lap  or  lead  at  one  end  than 
at  the  other.  When  a  link  is  interposed  between  the  valve  and  the  eccentric  as  shown 
in  the  locomotive  valve  gear,  Fig.  29,  then  the  valve  can  be  made  to  cut  off  equal 
portions  of  steam  in  the  forward  and  return  strokes,  without  making  a  diffei-ence  in  the 
lap  or  lead  of  the  valve.  Indeed,  in  locomotives  the  amount  of  lap  at  each  end  of  the 
valve  is  always  equal,  and  the  lead  for  full  stroke  at  each  end  of  the  valve  is  also  equal. 
88.  When  the  valve,  or  the  valve  gear,  is  constructed  so  as  to  make  the  valve  cut 
off  equal  portions  of  steam  in  the  forward  and  return  strokes,  the  cut-off  is  said  to  be 


Fig.82 


MODERX  LOCOMOTIVE   CONSTRUCTION.  73 

equalized.  To  equalize  the  cut-off  in  a  locomotive,  the  saddle-pin  A  is  moved  out  of 
the  center  of  the  link,  as  shown  in  Fig.  82,  that  is,  the  center  of  the  saddle-pin  is  moved 
a  certain  distance  towards  the  center  from  which  the  link  has  been  drawn;  and  besides 
this,  the  lifting-shaft  arms  must  be  made  of  the  proper  length,  and  the  lifting  shaft 
placed  in  the  correct  position.  To  determine  how  much  the  saddle-pin  must  be  moved 
out  of  center,  and  what  length  the  lifting-shaft  arms  should  be  made,  and  where  to 
place  the  lifting  shaft  so  that  the  cut-off  will  be  equalized,  we  shall  show  later. 

In  a  locomotive  it  is  of  great  importance  to  equalize  the  cut-off,  as  this  will  cause 
the  engine  to  work  smoother  and  better  than  with  a  cut-off  not  equalized.  Again,  an 
ei|uali/.e<l  cut-off  will  produce  an  exhaust  at  regular  and  equal  intervals,  an  attainment 
which  is  in  itself  of  the  greatest  importance,  because  when  an  engine  is  running  the 
sound  of  the  exhaust  indicates  to  the  engineer  the  working  conditions  of  such  parts  of 
mechanism  as  are  out  of  sight;  hence  the  engineer,  besides  keeping  a  strict  look-out 
for  the  parts  of  mechanism  which  are  in  sight,  and  performing  other  duties  imposed 
upon  him,  constantly  listens  to  the  exhaust,  and  as  long  as  this  beats  at  regular  and 
equal  intervals  he  knows  that  the  valve  and  valve  gear  are  in  good  working  order,  or, 
so  to  speak,  are  in  a  healthy  condition ;  but  as  soon  as  the  exhaust  commences  to  beat 
at  irregular  or  unequal  intervals,  the  engineer  accepts  this  fact  as  a  warning  that  some- 
thing is  seriously  wrong,  and  that  an  immediate  examination  of  his  engine  is  absolutely 
necessary. 

89.  In  the  foregoing  construction  we  have  assumed  that  the  eccentric-rod  is  of  an 
infinite  length.  Such  an  assumption  will  in  no  wise  interfere  with  the  positions  of  the 
points  y  and  ar,  which  indicate  the  position  of  the  center  of  eccentric  when  the  piston 
is  at  the  beginning  of  the  forward  and  return  strokes,  and  if  these  points  are  located 
with  absolute  exactness,  according  to  foregoing  instructions,  the  positions  of  these 
points  will  be  absolutely  correct.  But  these  remarks  do  not  apply  to  the  points  g  and 
g.2,  which  indicate  the  positions  of  the  center  of  eccentric  at  the  moment  that  the  steam 
is  cut  off  in  the  forward  and  return  strokes. 

To  find  the  correct  positions  of  the  points  g  and  gt,  we  should  take  into  considera- 
tion the  length  of  the  eccentric-rod,  which,  on  account  of  its  angularity  during  the 
travel  of  the  valve,  will  somewhat  affect  the  positions  of  the  points  //  and  g.2,  and  conse- 
quently the  points  of  cut-off  will  also  be  affected.  Yet  the  change  in  the  positions  of 
the  points  f/  and  //._,  which  will  occur  when  the  length  of  the  eccentric-rod  is  taken  into 
consideration,  will  be  so  slight,  that  for  ordinary  engines  it  will  hardly  be  appreciable, 
and  therefore  may  be  neglected.  But,  in  equalizing  the  cut-off  in  locomotives,  the 
length  of  the  eccentric-rod  is  taken  into  account,  as  will  be  seen  later. 


CHAPTER    III. 

VALVE  GEAR.— CONSTRUCTION   OF  LINKS. 

KOCKERS. 

90.  Figs.  83  and  84  represent  a  rocker  such  as  is  used  in  American  locomotives. 
The  general  practice  in  locomotive  construction  is  to  make  the  rockers  of  wrought-iron, 
although  occasionally  we  find  them  made  of  cast-iron.  The  arms  and  the  shaft  are 
forged  in  one  piece.  The  holes  in  the  end  of  the  arms  should  be  tapered,  the  taper 
being  equal  to  J  of  an  inch  in  12  inches ;  that  is,  in  a  hole  12  inches  deep  its  diameter 
at  one  end  should  be  £  of  an  inch  larger  than  the  diameter  at  the  other  end.  Some 
makers  adopt  a  greater  taper,  sometimes  as  great  as  3  of  an  inch  in  12  inches.  The 
pins  are  made  to  fit  these  holes  very  accurately,  so  that  they  must  be  driven  home 
with  a  hammer.  The  pins  are  made  of  wrought-iron,  and  are  case-hardened. 

The  reason  for  making  the  holes  in  the  rocker-arms  tapered,  is  that  the  pins  can 
be  driven  out  with  greater  ease,  and  without  injuring  or  upsetting  their  ends.  On  the 
other  hand,  if  these  holes  have  no  taper,  and  the  pins  are  fitted  into  the  rockers  as  firmly 
as  is  required  in  a  locomotive,  then  the  pins  will  need  so  much  hammering  in  driving 
them  into  or  out  of  the  holes  as  to  produce  injurious  result. 

The  design  of  the  locomotive  generally  limits  the  length  of  the  rocker-shaft  between 
the  arms,  but  when  there  is  room  enough  the  length  of  the  rocker-shaft  should  be 
at  least  12  inches  for  large  locomotives,  and  not  less  than  9  inches  for  smaller 
engines.  The  diameter  of  the  shaft  must  be  such  that  it  will  have  sufficient  strength  to 
resist  the  severest  stress  to  which  the  rocker  may  be  subjected  without  springing  or 
twisting  the  shaft  to  any  appreciable  extent. 

The  greatest  stress  to  which  a  rocker  can  be  subjected  will  occur  when  an  engine 
is  to  be  started  after  it  has  been  allowed  to  stand  still  for  some  time,  for  the  following 
reasons : 

1st.  When  an  engine  is  to  be  started  after  standing  still  for  some  time,  the  valve 
seat  will  be  dry,  and  it  will  not  be  lubricated  until  the  engine  has  made  a  few  turns, 
consequently  a  greater  force  will  be  required  to  move  the  valve  when  the  engine  is 
commencing  to  move  than  after  it  has  been  in  motion  for  some  time. 

2d.  Besides  the  force  necessary  to  overcome  the  friction,  an  additional  force  will 
be  required  to  overcome  the  adhesion  caused  by  the  oil  that  remained  between  the 
valve  and  its  seat  when  the  engine  was  stopped. 

3d.  In  starting  an  engine  the  valve  may  occupy  a  position  as  shown  in  Fig.  84, 
that  is,  the  valve  covering  both  steam  ports.  When  this  happens  the  friction  between 


MODERX    LOCOMOTIVE    COXSTRUCTION. 


75 


the  valve  and  its  seat  will  not  be  diminished  by  any  reaction  of  steam  pressure  in  the 
exhaust  or  steam  ports ;  therefore  the  friction  between  the  valve  and  its  seat  will  be 
proportional  to  the  total  steam  pressure  upon  the  back  of  the  valve.  Here,  then,  we 
see  that,  in  starting  an  engine,  considerably  more  power  will  be  required  to  work  the 
valve  tluin  will  be  necessary  to  work  the  same  valve  after  the  engine  has  been  in  motion 
for  some  time.  Again,  we  must  not  neglect  the  force  required  to  overcome  the  friction 
between  the  packing  in  the  stuffing-box  and  the  valve-stem,  which  at  times  may  be 
considerable,  caused  by  carelessly  tightening  the  stuffing-box  gland.  Thus  we  can 
understand  what  forces  a  rocker  has  to  overcome;  and  it  must  be  made  strong  enough 
to  do  it.  But  to  calculate  the  exact  amount  of  force  necessary  to  move  a  slide-valve 
uiiiler  the  foregoing  conditions  is  impossible;  we  can  only  adopt  an  empirical  rule, 
the  correctness  of  which  is  based  upon  close  observation  in  actual  practice.  The 
writer  believes  that,  by  making  a  rocker  strong  enough  to  overcome  £  of  the  total 
strain  pressure  on  the  back  of  the  valve,  good  results  will  be  obtained.  Thus,  for 
instance : 

EXAMPLE  27. — The  length  of  the  slide-valve  is  9  inches,  its  breadth  16  inches,  and 
the  steam  pressure  in  the  steam-chest  is  120  pounds  per  square  inch ;  what  will  be  the 
greatest  force  that  the  rocker  must  be  capable  of  overcoming? 

The  total  steam  pressure  upon  the  back  of  the  valve  is  obtained  by  multiplying 
the  area  of  the  valve  face  by  the  steam  pressure  per  square  inch;  hence  we  have 
9"  x  16"  x  120  =  17,280  pounds,  which  is  the  total  pressure  upon  the  back  of  the  valve, 
and  J  of  this  pressure  will  be  the  force  that  the  rocker  must  be  capable  of  over- 
coming without  twisting  or  springing  the  shaft ;  therefore  -^f^  =  5,760  pounds  will 


Fig.  86  I 


be  the  greatest  stress  to  which  the  rocker  can  be  subjected,  or,  in  other  words,  the 
greatest  force  that  it  must  overcome.  Now,  when  we  know  this  force  we  can  easily 
determine  by  computation  the  suitable  diameter  of  the  rocker-shaft. 

91.  We  find  in  books  relating  to  the  strength  of  material,  that  in  the  instance  of  a 
shaft  which  is  firmly  fixed  at  one  end,  having  a  lever  attached  to  its  other  end  with  a 
force  applied  to  the  end  of  the  lever,  its  diameter,  which  will  be  sufficiently  large  to  resist 
twisting,  is  determined  by  multiplying  the.  length  of  the  lever  in  inches  by  the  force  in 
pounds  applied  to  the  end  of  tlie  lever,  then  dividing  this  product  by  a  constant  quantity 
which  has  been  previously  obtained  by  actual  experiment,  and  extracting  the  cube  root 
of  the  quotient ;  the  result  will  be  the  diameter  of  the  shaft  in  inches. 


76  MODERN   LOCOMOTIVE    CONSTRUCTION. 

This  rule  is  used  for  determining  the  diameter  of  a  rocker-shaft,  and  when  we 
apply  it  we  must  assume  the  upper  rocker-shaft  arm  11,  Figs.  83  and  84,  to  be  the 
lever  attached  to  the  shaft,  its  length  being  the  distance  between  the  center  of  shaft 
and  the  center  of  pin ;  the  constant  quantity  (before  alluded  to)  may  be  taken  at 
1,200.  Hence,  to  find  the  diameter  of  a  rocker-shaft  made  of  wrought-iron,  we  have 
the  f ollowing : 

RULE  14. — Multiply  the  length  of  the  upper  rocker-arm  in  inches  by  &  of  the  total 
steam  pressure  upon  the  back  of  the  valve ;  divide  this  product  by  1,200,  and  the  cube 
root  of  the  quotient  will  be  the  diameter  of  the  shaft  in  inches ;  or  putting  this  rule  in 
the  shape  of  a  formula,  we  have 


/Length  of  rocker-arm  in  indies  x   '  stpilm  in-cssnvc  on  hack  of  valve. 
Diameter  of  shaft  =  \  /  - 

\f 

If  the  rocker-shaft  is  to  be  made  of  cast-iron,  then  the  same  rule  is  applicable,  with 
this  exception,  instead  of  using  the  constant  quantity  1,200,  we  must  use  the  constant 
1,000. 

EXAMPLE  28. — Find  the  diameter  of  a  wrought-iron  rocker-shaft  which  has  to  move 
a  slide-valve  9  inches  long  and  16  inches  wide ;  the  steam  pressure  in  the  steam-chest 
is  120  pounds  per  square  inch,  and  the  length  of  the  upper  rocker-arm  10  inches. 

We  have  seen  in  Example  27  that  £  of  the  total  steam  pressure  on  the  back  of  the 
valve  9  inches  long  and  16  inches  wide  is  5,760  pounds.  This  5,760  pounds  is  the  force 
applied  to  the  end  of  the  upper  rocker-arm ;  therefore,  according  to  Rule  14,  we  have, 

10"  x  5760  pounds  _ 
1200 

and  the  cube  root  of  48  is  3.63,  consequently  the  shaft  must  be  3|  (nearly)  inches  in 
diameter.  To  find  the  cube  root  in  an  easy  manner  of  any  quantity,  we  simply 
refer  to  a  table  of  cube  roots,  which  will  be  found  in  any  good  engineer's  pocket-book. 

92.  The  length  of  the  upper  rocker-arm  »,  Figs.  83  and  84,  is  limited  in  either 
direction ;  that  is,  it  must  not  be  made  too  long  or  too  short.  For  a  slide-valve  having 
5  inches  travel,  the  length  of  the  upper  arm  is  generally  10  inches,  and  for  valves 
with  less  travel,  the  length  of  the  upper  arm  can  be,  and  should  be,  somewhat  reduced. 
The  reason  for  this  is:  If  the  length  of  the  upper  rocker-arm  for  a  valve  having 
5  inches  travel  is  made  much  longer  than  10  inches,  then  the  shaft  will  be  subjected  to 
a  greater  twisting  stress,  and  consequently  the  diameter  of  the  shaft  must  be  increased, 
which  is  not  always  desirable.  The  reason  why  the  diameter  of  the  shaft  should  be 
made  larger  when  the  upper  rocker-arm  is  made  longer  can  easily  be  seen  by  examin- 
ing Rule  14. 

Again,  this  arm  should  not  be  made  much  shorter  than  10  inches  for  a  valve 
with  5  inches  travel,  on  account  of  the  custom  that  is  followed  by  the  majority  of 
locomotive-builders  and  master-mechanics  of  keying  the  valve-rod  to  the  valve-stem,  as 
shown  in  Fig.  84,  thus  making  a  rigid  connection.  Now  notice,  in  Fig.  84,  that  the 
path  of  the  valve-rod  pin  is  an  arc,  as  x  y,  and  to  this  arc  the  end  of  the  valve-rod 
must  accommodate  itself ;  and  since  the  valve-stem  must  travel  in  a  straight  line,  and 
since  there  is  not  a  flexible  joint  between  the  valve  rod  and  stem,  it  follows  that  the 
valve-rod  must  be  sprung  out  of  a  straight  line  during  the  travel  of  the  valve,  and  the 


MODERX   LOCO.VOTirE    COXSTRUCTIOX.  77 

amount  that  the  latter  is  sprung  out  of  a  straight  line  is  equal  to  the  line  a  ?>,  Fig.  84. 
Nmv,  it'  tin-  ui i] XT  rocker-arm  is  made  much  shorter  than  10  inches,  leaving  the 
tnivel'ot'  tlit-  valve  the  same  as  before,  then  the  line  a  l>  will  be  longer,  and  consequently 
tin-  amount  that  the  valve-rod  must  be  sprung  out  of  a  straight  line  during  the  travel 
of  the  valve  will  also  be  greater  than  before,  producing  injurious  results. 

The  width  of  the  rocker-arms  on  the  line  c  d,  passing  through  the  center  of  th6 
shaft,  as  shown  in  Fig.  84,  is  not  made  the  same  by  the  different  locomotive-builders  or 
master-mechanics,  yet  the  following  arbitrary  rule  will  give  a  width  agreeing  very 
closely  with  the  present  practice: 

RULE  15. — To  the  diameter  of  the  rocker-shaft  add  one-half  of  the  same  diameter, 
and  from  this  sum  subtract  £  of  an  inch.  Or,  in  the  shape  of  a  formula  we  have : 

Diameter  of  shaft  +  J  of  diameter  of  shaft  —  i  of  an  inch  =  width  c  d  of  rocker-arm. 

EXAMPLE  29. — Diameter  of  the  rocker-shaft  is  3£  inches ;  what  must  be  the  width 

of  the  rocker-arm  at  c  dl 

3f"  +  Hi"  -  |"  =  5fV',  say  5|". 

When  the  width  c  d  of  the  rocker-arm  is  known,  its  thickness  can  be  easily  ascer- 
tained  in  the  following  manner:  Assume  that  the  arm  is  a  lever  firmly  fixed  atone 
•  •iid  and  loaded  at  the  other  end,  as  shown  in  Fig.  85.  Then,  according  to  rules  found 
in  books  relating  to  the  strength  of  material,  we  find  that  the  load  which  a  beam  or 
lever,  firmly  fixed  at  one  end  and  loaded  at  the  other,  can  support  with  safety  is  deter- 
mined by  multiplying  the  square  of  the  width  c  d  by  the  thickness,  and  by  a  constant 
quantity,  previously  found  by  experiment  (this  quantity  is  generally  called  the  "  co- 
efficient "),  and  dividing  this  product  by  the  length  of  the  beam  or  lever.  In  applying 
this  rule  to  the  rocker  we  shall  adopt  1,200  for  the  constant  quantity  or  coefficient, 
and  the  dimensions  of  the  rocker  will  be  taken  in  inches ;  therefore  we  have, 

Square  of  the  width  c  d  x  thickness  x  1200 

,,    .     . — r-  -  =  load. 

Length  in  inches. 

Now,  if  we  know  the  load,  the  width  at  c  d,  and  the  length  of  the  rocker-arm,  but 
not  its  thickness,  we  can  establish  a  rule  from  the  foregoing  formula  which  will  enable 
us  to  find  the  thickness  of  the  rocker-arm. 

RULE  16. — Multiply  the  load  in  pounds  by  the  length  of  the  rocker-arm  in  inches, 
and  divide  this  product  by  the  square  of  the  width  in  inches  into  1,200;  the  quotient 
will  be  the  thickness  of  the  arm.  Or,  in  the  shape  of  a  formula  we  have, 

Load  x  length  in  inches 


Square  of  the  width  c  d  in  inches  x  1200 


=  thickness  in  inches. 


EXAMPLE  30. — What  must  be  the  thickness  of  the  upper  rocker-arm,  its  width  c  d 
lidng  .~>i  inches;  length,  1(1  indies;  and  the  valve  which  the  rocker  has  to  move  is 
9"  x  16";  the  steam  pressure  in  the  steam-chest  is  120  pounds? 

The  total  load  which  a  rocker  has  to  support  is  equal  to  the  greatest  stress  to  which 
it  can  be  subjected,  and.  as  we  have  seen  before,  the  latter  is  equal  to  J  of  the  total 
steam  pressure  upon  the  back  of  the  valve. 


78 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


Now,  according  to  Example  27,  we  know  that  £  of  the  total  steam  pressure  upon 
the  back  of  the  valve  9"  x  16"  and  with  a  pressure  of  120  pounds  per  square  inch  is 
equal  to  5,760  pounds,  which  must  now  be  considered  as  the  load ;  consequently  we 
have,  according  to  Rule  16, 

5760  x  10" 

Thickness  =  -  -  =  1.58  or  1A"  inch  nearly, 

(square  of  5£)  x  1200 

which  is  the  thickness  of  the  rocker-arm  without  the  hub. 

93.  In  a  great  many  locomotives  we  find  the  hub  h  on  the  upper  rocker-arm  placed 
upon  the  outside  of  the  latter,  as  shown  in  Fig.  86.  This  should  be  avoided  as  much  as 
possible,  because  when  the  hub  is  so  placed  the  arm  will  not  only  have  to  resist  a  trans- 
verse stress,  but  also  an  increased  twisting  stress,  and  therefore  the  arm  and  shaft  must 
be  made  correspondingly  strong.  If  the  hub  is  placed  upon  the  inside  of  the  rocker- 
arm,  as  shown  in  Fig.  83,  then  the  twisting  stress  will  be  reduced,  but  yet  not  alto- 
gether removed.  To  allow  for  this  extra  twist- 
ing stress  which  still  remains,  we  have  adopted 
in  the  foregoing  rules  a  coefficient  of  1,200.  If 
the  rocker-arm  had  no  twisting  stress  to  resist, 


H 


•• 
' 

^v«t 

5 

FifJ.89 

[ 

/,-,,<*  ml  lo',Cy.ls 

DJ 


bat  simply  a  transverse  stress,  then  the  coefficient  of  1,200  could  have  been  increased 
to  1,800 ;  the  result  of  this  would  have  been  a  thinner  rocker-arm. 

On  the  lower  rocker-arm  we  are  generally  compelled  to  place  the  hub  on  the  outside 
of  the  arm,  because  room  is  required  to  clear  the  eccentric-rod  jaw  and  pin,  as  shown  in 
Fig.  83.  Figs.  87, 88, 89, 90,  and  91  show  the  wrought-iron  rockers  for  the  different  sizes 
of  locomotives.  The  diameters  of  the  shafts  and  dimensions  of  rocker-arms  have  been 
obtained  according  to  the  foregoing  rules,  and  are  suitable  for  cylinders  whose  ports 


MODERX  LOCOMOTIVE   COXSTRVCTION. 


79 


are  proportioned  for  a  piston  speed  of  600  feet  per  minute,  the  valves  having  the  ordi- 
nary  amount  of  lap,  with  a  steam  pressure  120  pounds  per  square  inch  in  the  steam- 
chest.  Comparing  the  dimensions  given  in  these  illustrations  with  the  dimensions  of 
rockers  in  actual  use  in  modern  locomotives,  it  will  be  seen  that  the  former  agree  very 
closely  with  the  latter.  Of  course,  the  length  of  arms  given  in  the  illustrations  may 
have  to  be  changed  to  suit  some  particular  design  of  engine,  and  when  the  change  is 
very  great,  then  the  dimensions  of  arms  and  shaft  should  be  determined  according  to 
the  foregoing  rules. 

Again,  to  be  very  exact,  we  should  have  given  a  differently  proportioned  rocker 
for  each  size  of  cylinder,  but  this  would  cause  a  complication  of  patterns,  which 
managers  of  private  establishments  and  master-mechanics  on  railroads  seek  to  avoid ; 
hence  one  size  of  rocker  is  generally  used  for  two  or  three 
different  sizes  of  cylinders — that  is,  cylinders  varying  in 
diameter. 

Now,  as  simple  as  a  rocker  may  appear  to  an  ordinary 
observer,  it  requires  care  to  proportion  it.  If  the  rocker 
is  made  too  weak,  it  may  still  be  strong  enough  to  move 
the  valve,  yet  it  will  spring  sufficiently  to  derange  the 
whole  valve  motion.  Indeed,  we  have  met  with  locomo- 
tives  which,  on  leaving  the  round-house,  had  such  an 
irregular  exhaust  that  the  engineers  stopped  the  engines 
and  examined  the  valve  motions,  but  found  that  the 
cause  of  all  the  trouble  was  the  springing  of  the  rocker- 
shaft — a  trouble  which  would  disappear  after  the  valve 
face  and  seat  became  lubricated  by  running  the  engines  a 
short  distance. 

94.  When  the  valve-rod  is  comparatively  very  short, 
a  knuckle-joint  must  be  introduced  between  valve  rod 
and  stem,  as  illustrated  in  Figs.  92,  93,  and  94,  and  which  needs  no  further  explanation. 

In  a  few  cases  the  valve-rod  end  which  connects  to  the  rocker-pin  is  provided 
with  brasses  and  a  key,  so  as  to  take  up  the  wear.  But  the  almost  universal  practice 
in  the  construction  of  American  locomotives  is  to  drive  a  bush  into  the  eye  of  the 
valve-rod,  as  shown  in  Fig.  95  (page  75).  This  bush  is  made  of  wrought-iron,  gen- 
erally ^  of  an  inch  thick  for  large  locomotives  and  i  of  an  inch  for  smaller  en- 
gines. The  bush  is  bored  and  turned,  and  then  case-hardened,  and  finally  forced 
into  the  eye  of  the  valve-rod  by  an  hydraulic  press.  Valve-rods  with  case-hardened 
bushes  will  need  but  very  little  repair,  as  the  wear  is  comparatively  slow,  and  when 
the  wear  of  the  bush  becomes  too  great,  it  can  be  easily  removed  and  replaced  by 
a  new  one  with  very  little  expense. 


ftg.92 


Talve  rod 


ECCENTRICS   AND   STRAPS. 


95.  In  the  following  illustration  we  have  represented  various  eccentrics  and  their 
si  nips,  with  all  the  important  dimensions  marked  upon  them.  These  have  been  selected 
from  a  number  of  designs  adopted  by  some  of  our  best  locomotive  builders.  Figs. 


80 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


Jflg-  147 


Fly.  *4S 


ftg,  J-40 


143  and  144  represent  two  views  of  an  eccentric,  Figs.  145  and  146  represent  two 
views  of  its  strap,  and  Fig.  147  a  section  of  the  same.     Such  eccentrics  and  straps  are 

used  on  some  of  our  large  loco- 
motives, that  is,  consolidation 
engines  having  cylinders  20"  in 
diameter  and  24"  stroke.  Figs. 
152  and  153  represent  an  eccen- 
tric, and  Figs.  154,  155,  156  rep- 
resent the  strap  for  the  same. 
Eccentrics  and  straps  of  this  size 
are  used  on  our  smallest  loco- 
fiff.  140  motives,  namely,  eight-wheeled 

w 


passenger  engines,  with  cylinders 
10"  in  diameter  and  18"  stroke. 
When  we  say  "  smallest  locomo- 
tives" we  do  not  include  loco- 
motives for  mining  purposes,  or 
very  light  narrow-gauge  loco- 
motives. 

The  duty  of  an  eccentric  and 
its  strap  is  to  move  the  slide- 
valve  forward  and  backward; 
and  when,  a  few  years  ago, 
pumps  were  used  in  locomotive 
engines,  then  occasionally  an 
eccentric  was  employed  to  work 
the  pump.  The  action  of  an  ec- 
centric, as  we  have  stated  in 
Art.  55,  is  precisely  the  same  as 
that  of  a  crank.  No  peculiar 
movement  must  be  expected  by 
the  use  of  an  eccentric  ;  the 
slide-valve  will  perform  its  func- 
tions as  correctly  when  it  re- 
ceives its  movement  from  a  crank 
as  when  it  receives  its  move- 
ment from  an  eccentric  ;  the 
only  reason  why  an  eccentric  is 
adopted  is,  that  the  use  of  the 
crank  is  impracticable.  In  Amer- 
ican locomotives  the  eccentrics 
and  straps  are  generally  made  of 
cast-iron ;  indeed,  we  may  say  they  are  always  made  of  cast-ii'on,  as  we  seldom  find 
a  locomotive  whose  eccentrics  and  straps  are  made  of  brass  or  of  wrought-iron. 
To  prevent  the  strap  from  slipping  sideways  off  the  eccentric,  a  recess  marked  (7, 


Flg.157 


v<n>t:  i;\  LocoMorirE  coxsTRUcnoy.  81 

Fig.  147,  is  turned  in  the  strap,  which  fits  a  corresponding  projection  turned  on  the 
cccciitric. 

Some  builders  make  the  joint  K  L,  Fig.  145,  perpendicular  to  the  center  line  M  N 
of  the  eccentric-rod ;  others  make  this  joint  not  at  right  angles  to  the  center  line  M  N, 
as  shown  in  Fig.  157.  The  advantage  claimed  for  the  latter  is  that  the  stress  will  be 
less  on  the  nuts  of  the  bolts  which  hold  the  two  parts  of  the  strap  together.  The  advan- 
tage claimed  for  the  former  design  is  that  no  right-  and  left-hand  pattern  for  the  strap 
will  be  required.  The  oil-cup  is  screwed  in  one  of  the  hubs  J,  Fig.  145 ;  the  reason  why 
two  hubs  are  cast  on  the  strap  is,  as  before,  to  avoid  a  right-  and  left-hand  pattern. 

The  eccentric-rod  fits  into  the  recess  marked  E  E.>,  Figs.  145  and  147,  and  is  secured 
to  the  strap  by  three  bolts.  It  will  be  noticed  in  Fig.  145  that  the  hole  for  one  bolt  is 
oblong ;  this  will  allow  the  rod  to  be  moved  outward  or  inward  in  the  recess,  as  may  be 
required,  and  then  fastened.  After  the  correct  position  of  the  rod  in  the  recess  has  been 
found,  then  the  other  two  holes  are  drilled,  reamed,  and  the  bolts  driven  in  tight,  so  that 
tin'  distance  from  the  center  of  the  strap  to  the  extreme  end  of  the  rod  cannot  be  changed. 
Some  master-mechanics  object  to  this  arrangement,  and  prefer  to  let  the  eccentric-rod 
butt  against  the  eccentric-strap,  as  shown  in  Fig.  149.  In  this  case,  the  distance 
between  the  center  of  the  strap  and  the  extreme  end  of  the  rod  can  be  changed  by  plac- 
ing some  thin  copper  strips  between  the  strap  and  the  rod. 

The  eccentric  is  generally  cast  in  one  piece,  but  sometimes,  for  the  sake  of  conven- 
ience in  repairing,  it  is  made  in  two  parts,  as  shown  in  Figs.  148  and  151.  For  holding 
the  two  parts  of  the  eccentric  firmly  together,  some  master-mechanics  use  studs  and 
nuts,  as  shown  in  Fig.  148 ;  others  use  studs  with  split  keys  or  cotters,  as  shown  in  Fig. 
151.  The  writer  believes  that  the  latter  method  is  the  best,  since  for  the  want  of  room 
in  the  design  shown  in  Fig.  148  it  is  often  extremely  difficult  to  gain  access  with  a 
wrench  to  the  nuts. 

During  the  time  of  setting  the  valve  gear  the  eccentrics  are  held  in  position  by  the 
set-screws,  but  afterwards,  in  the  majority  of  locomotives,  they  are  keyed  to  the  axle. 
Of  course,  in  a  small  number  of  locomotives,  as  may  be  inferred  from  the  foregoing 
remark,  the  eccentrics  are  not  keyed  to  the  axle,  and  are  held  in  position  by  the  set 
screws  only.  The  set-screws  are  made  of  steel,  cupped  as  shown  in  Fig.  159,  and  then 
hardened.  The  key-way  in  the  eccentric  is  cut  before  the  latter  is  placed  on  the  axle, 
but  the  key-way  in  the  axle  is  cut  after  the  correct  position  of  the  eccentric  has  been 
found.  To  cut  this  key- way — which  must  be  done  by  hand — is  very  troublesome; 
hence  some  master-mechanics  cut  no  key-ways  in  the  axle,  but  use  two  keys,  as  shown 
in  Fig.  160.  In  this  case  each  key  has  teeth  cut  lengthways  on  one  of  its  sides,  as 
shown,  so  that  when  the  keys  are  driven  home  the  teeth  will  grip  the  axle. 

The  set-screws  in  this  case  are  used  as  an  extra  security  against  the  slipping  of  the 
keys  out  of  position. 

The  form  of  the  section  of  the  strap  can  be  made  as  shown  in  Figs.  150  and  156,  or 
as  represented  in  Fig.  158.  By  adopting  the  form  shown  in  the  latter  figure,  the  strap 
can  be  made  lighter,  and  still  have  the  same  strength  as  that  shown  in  Fig.  156,  but  the 
outside  diameter  of  the  strap  in  Fig.  158  will  necessarily  be  somewhat  larger  than  that 
in  Fiir.  15C. 

'.Hi.  The  diameter  of  the  eccentric,  and  also  that  of  the  strap,  should  be  made  as 


82 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


Fig. 142 


small  as  possible,  since  by  so  doing  not  only  material  will  be  saved,  but  also  the  work 
expended  in  friction  will  be  reduced ;  and  lastly,  less  space  will  be  required  for  the 
eccentric  to  work  in.  This  last  fact  is  of  great  importance,  because  in  a  great  number 
of  locomotives  the  space  required  for  the  revolving  of  the  eccentrics  limits  the  length 
of  the  fire-box ;  and  since  the  available  space  for  the  fire-box  is  often  barely  sufficient, 
it  follows  that  space  should  not  be  wasted  by  making  the  eccentrics  larger  than  neces- 
sary. Thus,  for  instance,  in  eight-wheeled  passenger  engines  the  main  driving-axle  on 

which  the  eccentrics  are  fastened  is  placed  compara- 
tively close  to  the  front  end  of  the  fire-box,  as  shown 
in  Fig.  142,  and  this  distance  between  the  main  axle 
and  the  fire-box  is  determined  by  the  space  required 
for  the  working  of  the  eccentrics.  Now,  whatever 
amount  the  radius  of  the  eccentric  is  made  too 
large,  that  same  amount  must  be  taken  from  the 
length  of  the  fire-box,  thus,  to  some  extent,  reduc- 
ing the  steaming  capacity  of  the  engine. 

Here,  then,  we  see  the  necessity  of  making  the 
diameter  of  the  eccentric  and  that  of  its  strap  as 
small  as  possible.  The  distance  between  the  cen- 
ter of  the  main  axle  and  that  of  the  fire-box  is 
generally  14  to  14 \  inches  in  large  locomotives, 
and  from  10  to  10J  inches  in  the  smaller  locomo- 
tives. But  here,  then,  the  question  arises,  How  can 
the  correct  diameter  of  the  eccentric  be  determined  ?  In  order  to  find  the  diameter  of 
an  eccentric,  we  must  know  its  eccentricity ;  that  is,  the  distance  between  the  center  y 
of  the  axle  and  the  center  x  of  the  eccentric,  Fig.  142.  We  must  also  know  the  diame- 
ter of  the  axle  on  which  the  eccentric  is  fastened,  and  the  thickness  of  the  metal  at  C. 
When  these  items  are  known,  we  add  together  the  distance  between 
the  centers  x  and  y,  half  the  diameter  of  the  axle,  and  the  thickness 
of  the  metal  at  (7,  and  multiply  this  sum  by  2.  Thus,  for  instance, 
in  Fig.  143  (Art.  95)  we  see  that  the  distance  between  the  center 
of  axle  and  the  center  of  eccentric — that  is,  the  eccentricity — is 
2£  inches,  half  the  diameter  of  the  axle  is  3£  inches,  and  the  thick- 
ness of  metal  at  C  is  l£  inches.  Adding  these  dimensions  together, 
we  have  2^  +  3£  +  l£  =  7£,  and  7£  x  2  =  15  inches,  which  is  the 
diameter  of  the  eccentric.  From  this  we  see  that  there  are  three 
items  whose  dimensions  determine  the  diameter  of  the  eccentric, 
namely,  its  eccentricity,  the  thickness  of  metal  at  (7,  and  the  diam- 
eter of  the  axle.  Here,  then,  another  question  arises :  How  can  we  find  the  dimen- 
sions of  these  three  items  ?  The  diameter  of  the  axle  is  determined  by  rules  to  be 
explained  hereafter ;  hence  there  remain  only  the  two  former  items,  whose  dimensions 
will  claim  our  consideration. 

97.  In  Art.  55  we  have  stated  that  the  throw  of  an  eccentric  is  equal  to  twice  its 
eccentricity,  hence  the  throw  of  the  eccentric  shown  in  Fig.  143  will  be  5  inches ;  we 
have  also  stated  that  the  throw  is  equal  to  the  travel  of  the  valve  for  engines  in  which 


d 

(ft 

\A  \ 

V 

I 

h 

& 

Fig.  101 


xonf:n\  locoMonrE  COSSTRITTION. 


83 

no  rocker  is  interposed.  We  will  now  add  to  this  statement  that,  when  a  rocker  is 
used  whose  arms  are  of  equal  lengths,  then  the  throw  of  an  eccentric  will  still  be  equal 
to  the  travel  of  the  valve ;  on  the  other  hand,  if  a  rocker  is  used  whose  arms  are  not  of 
equal  lengths,  then  the  throw  will  not  be  equal  to  the  travel  of  the  valve.  Conse- 
quently, when  the  travel  of  the  valve  is  given  for  an  engine  that  has  no  rocker,  or  when 
the  travel  of  a  valve  is  given  for  an  engine  in  which  a  rocker  is  employed  whose  arms 
are  of  equal  length,  in  either  case  we  must  make  the  eccentricity  of  the  eccentric  equal 
to  one- half  of  the  travel  of  the  valve ;  we  cannot  make  it  less  or  more ;  hence,  in  these 
two  cases  the  shoiiest  distance  between  the  centers  x  and  y,  Fig.  142,  will  be  equal  to 
half  the  travel  of  the  valve. 

When  a  rocker  is  used  whose  arms  are  not  of  equal  lengths,  then  the  eccentricity 
of  the  eccentric  will  be  either  more  or  less  than  half  the  travel  of  the  valve. 

Thus  for  instance : 

EXAMPLE  31. — Fig.  161.  If  the  upper  arm  A  of  the  rocker  is  10  inches,  and  the 
lower  arm  B  is  12  inches  long,  and  the  travel  of  the  valve  5  inches,  what  will  be  the 
eccentricity  of  the  eccentric  ? 

•  Let  the  line/0  represent  the  position  of  the  center  of  the  rocker  when  the  valve 
stands  midway  of  its  travel ;  from  the  center  c,  and  with  a  radius  of  10  inches,  describe 
the  arc  d  e;  on  this  arc  lay  off  a  point  d  2£  inches  (one-half  of  the  travel)  from  the  cen- 
ter line/0,  not  measured  on  the  arc,  but  on  a  straight  line  perpendicular  to/0;  in  a 
similar  manner  lay  off  the  point  e  2%  inches  from/0;  then  the  point  d  will  represent 
the  position  of  the  center  of  the  rocker-pin  when  the  valve  stands  at  one  end  of  its 
travel,  and  the  point  e  will  represent  the  position  of  the  center  of  the  rocker-pin  when 
the  valve  stands  at  the  other  end  of  its  travel.  From  the  point  c  as  a  center,  and  with 
a  radius  of  12  inches,  describe  the  arc  h  i ;  through  the  point  d  and  the  center  c  draw  a 
straight  line  intersecting  the  arc  /*  i  in  the  point  * ;  also  through  the  point  e  and  the 
center  c  draw  another  straight  line  intersecting  the  arc  h  i  in  the  point  h.  The  distance 
between  the  points  h  and  i  will  be  equal  to  the  throw  of  the  eccentric,  and  half  of  this 
distance  will  be  the  eccentricity  of  the  eccentric.  If  this  drawing  is  accurately  made 
it  will  be  found  that  the  throw  is  6  inches,  hence,  in  this  case,  3  inches  will  be  the  dis- 
tance between  the  centers  x  and  y  in  Fig.  142,  and  is  £  inch  more  than  half  the  travel 
of  the  valve. 

EXAMPLE  32. — But  now  suppose  the  upper  arm  A  of  the  rocker  is  12  inches  long, 
and  the  length  of  the  lower  arm  B,  10  inches,  and  the  travel  of  the  valve  5  inches  as 
before,  then  what  will  be  the  eccentricity  of  the  eccentric  ? 

Fig.  162.  From  the  center  c,  and  with  a  radius  of  12  inches,  describe  arc  d  e ;  on 
this  arc  lay  off  as  before  points  d  and  c,  each  point  being  placed  2£  inches  from  the 
center  line  fg ;  then  the  distance  between  these  two  points  will  be  equal  to  the  travel 
of  the  valve.  From  c  as  a  center,  and  with  a  radius  of  10  inches,  describe  the  arc  h  i ; 
through  the  point  d  and  the  center  c  draw  a  straight  line  intersecting  the  arc  h  i  in  the 
point  /,  also  through  the  point  e  and  the  center  c  draw  a  straight  line  intersecting  the 
arc  h  i  in  the  point  Ji ;  the  distance  between  the  points  h  and  i  will  be  equal  to  the  throw 
of  the  eccentric,  and  half  of  this  distance  will  be  the  eccentricity  of  the  eccentric.  If 
tliis  drawing  is  correctly  made,  it  will  bo  found  that  the  distance  between  the  points  d 
and  i — that  is,  the  throw  of  the  eccentric — is  4.16  inches,  consequently  the  eccentricity 


84  MODERN  LOCOMOTITK   CONSTRUCTION. 

of  the  eccentric  will  be  2.08  inches,  say  2  inches,  ^  of  an  inch  less  than  the  travel  of  the 
valve. 

The  throw  of  an  eccentric  in  the  last  two  examples  can  also  be  found  by  the 
"  simple  rule  of  three,"  or,  as  it  is  sometimes  called,  "  the  simple  rule  of  proportion." 
Thus,  take  Example  31 ;  instead  of  finding  the  throw  graphically  as  shown,  we  may 
find  it  thus : 

10"  :  12"  : :  5" : throw 


10)60 

6  inches  =  the  throw. 
Or,  if  we  take  Example  32,  we  have, 

12":  10":   :  5":  throw 
5 

12)50 

4.166  inches  =  throw. 

PROPORTIONS   OF  ECCENTRICS. 

98.  Table  12  gives  the  proportional  dimensions  of  the  important  parts  of  the 
eccentric  and  strap.  For  instance,  this  table  indicates  that  to  find  the  thickness  at  C, 
Fig.  142,  we  multiply  a  given  unit  by  1,  and  thus  obtain  the  dimension  at  C  in  inches. 
By  "  unit "  is  meant  a  certain  number  regarded  as  one,  so  that  when  this  unit  is  multi- 
plied by  the  numbers  as  indicated,  the  important  dimensions  in  inches  of  an  eccentric 
and  strap  will  have  been  obtained.  This  unit  vis  found  in  the  following  manner :  We 
may  assume  that  the  friction  which  the  eccentric  has  to  overcome  is  proportional  to 
the  total  steam  pressure  on  the  back  of  the  valve,  which  is  equal  (Art.  82)  to  the  area 
of  the  valve  face  multiplied  by  steam  pressure  per  square  inch,  consequently,  for  finding 
the  unit  we  have  the  following  empirical  rule : 

Multiply  the  square  root  of  the  total  pressure  on  the  back  of  the  valve  by  the 
decimal  .01,  the  product  will  be  the  unit  required ;  or,  putting  this  rule  in  the  shape  of 
a  formula,  we  have, 

.01  v'total  pressure  on  the  back  of  the  valve. 

Here  the  decimal  .01  is  arbitrary,  and  should  only  be  used  in  locomotive  practice, 
in  which  it  always  remains  the  same,  no  matter  whether  the  locomotive  is  large  or 
small.  Again,  notice  that  the  total  pressure  on  the  back  of  the  valve  depends  upon  the 
size  of  the  valve  face  and  the  steam  pressure  per  square  inch ;  and  since  the  sizes  of 
the  valve  faces  vary  in  the  different  locomotives,  it  follows  that  this  unit  in  Table  12 
will  also  vary  for  the  different  classes  of  engines. 

EXAMPLE  33. — Take,  for  example,  a  consolidation  engine  with  cylinders  20  inches 
in  diameter ;  the  average  size  of  the  valve  face  for  these  engines  is  10  inches  long  and 
20  inches  wide,  hence  the  area  of  the  valve  face  is  10"  x  20"  =  200  square  inches. 
Assume  that  the  steam  pressure  in  the  steam-chest  is  120  pounds  per  square  inch, 


MODERN  LOCOMOTIVE   CONSTRUCTION.  g5 

we  have  120  x  200  =  24,000  pounds,  which  is  the  total  pressure  on  the  back  of  the 
valve.  The  square  root  of  24,000  is  154  (here  the.  fraction  has  been  neglected),  and 
134  x.Ol  =  1.34,  which  is  the  unit  required.  If  now  we  multiply  this  unit  1.54  by  the 
numbers  given  in  Table  12,  we  shall  obtain  the  following  dimensions  of  an  eccentric 
and  strap  suitable  to  work  a  slide-valve  with  a  total  pressure  of  24,000  pounds  upon 
its  back.  Thus  (see  Fig.  142) : 

TABLE  12. 

A    =  Unit  x  1 
B   =  Unit  x  2.25 
C    =  Unit  x  1 
D   =  Unit  x  1.75 
E    =  Unit  x  2.3 
E2  =  Unit  x    .7 
F   =  Unit  x  2 

The  dimensions  at  A  will  be  1.54  x  1       =  1.54  inches. 

"  "  B  "  "  1.54  x  2.25  =  3.46  " 

"            "  "  C  "  "  1.54  x  1       =  1.54  " 

"            "  "  D  "  "  1.54  x  1.75  =  2.69  " 

"            "  "  E  "  "  1.54  x  2.3    =  3.54  " 

"            "  "  E.,  "  "  1.54  x    .7    =  1.07  " 

"            "  "  F  "  "  1.54  x  2 .     =  3.08  " 

EXAMPLE  34. — Now  take  a  small  eight-wheeled  passenger  engine,  with  cylinders  10 
inches  in  diameter.  The  average  size  of  the  valve  face  for  this  class  of  engines  is  6 
inches  long  and  ll£  inches  wide.  Again,  assume  that  the  steam  pressure  per  square 
inch  in  the  steam-chest  is  120  pounds.  In  this  case  we  have  6"  x  11.5"  x  120  =  8280 
pounds  pressure  on  the  back  of  the  valve,  and  .01 V8280  (that  is,  the  square  root  of 
8280  x  .01)  =.91,  which  is  the  unit  required.  Consequently,  the  dimensions  of  an 
eccentric  and  strap  suitable  to  work  a  valve  with  8,280  pounds  pressure  upon  its  back 
will  be  (see  Fig.  142) : 

The  dimensions  at  A  =  .91  x  1       =    .91  inches. 

"  "  "  B  =  .91  x  2.25  =  2.04  " 

"  "  "  C  =  .91  x  1      =    .91  " 

"  "  "  D  =  .91  x  1.75  =  1.59  " 

"  "  "  E  =.91  x  2.3    =2.09  " 

"  "  "  E,  =  .91  x    .7    =    .63  " 

"  "  "  F"  =  .91  x  2      =  1.82  " 

If  we  now  compare  the  dimensions  found  in  Example  33  with  the  dimensions  ob- 
tained by  measurements  of  eccentrics  in  use  as  shown  in  Figs.  143  and  145,  we  find  these 
to  agree  very  closely.  The  greatest  difference  between  any  two  dimensions  is  that  of 
the  breadth  of  the  strap  at  B.  In  our  illustration  the  breadth  is  J  of  an  inch  greater 
than  that  obtained  by  the  rule,  but  it  must  be  remembered  that  the  eccentric  and  strap 
shown  in  Figs.  143  and  145,  although  frequently  used  in  modern  locomotives,  is  very 


86 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


heavy  in  comparison  with  those  employed  in  a  great  many  other  locomotives.  We 
also  find  that  the  dimensions  found  in  Example  34  agree  closely  with  those  shown  in 
Figs.  152  and  154.  Here  the  greatest  difference  between  any  two  dimensions  is  that  of 
the  width  of  the  recess  E  for  the  eccentric-rod.  The  writer  believes  that  if  the  width 
of  the  recess  is  made  according  to  the  rule  given,  namely,  2f  inches  instead  of  2^  inches, 
good  results  will  follow. 

Lastly,  to  those  who  are  acquainted  with  locomotive  work,  it  may  appear  that 
depth  of  the  lug  at  F  is  very  great  when  compared  with  the  lugs  on  ordinary  eccentric 
straps,  but  in  the  writer's  opinion  this  is  a  great  improvement,  because  when  the  holes 
in  these  lugs  are  reamed,  the  bolts  turned  and  fitted,  so  that  they  must  be  driven  into 
position,  this  increased  depth  of  lug  will  to  a  great  extent  prevent  the  strap  from 
springing  out  of  its  true  circular  form. 


LINK  MOTION. 


99.  In  Art.  61  we  have  seen  that,  when  a  direct  connection  is  made  between  the 
eccentric  and  valve  (that  is,  when  no  rocker  is  employed),  as  shown  in  Fig.  163,  the 


eccentric  will  always  travel  ahead  of  the  crank.  Consequently,  if,  as  in  Fig.  163,  the 
crank-pin  occupies  the  position  A  as  shown,  and  is  to  rotate  in  the  direction  as  indi- 
cated by  the  arrow  marked  1,  then  the  position  occupied  by  the  eccentric  will  be  as 
shown  in  full  lines  and  with  its  center  at  B.  If,  on  the  other  hand,  the  crank-pin  oc- 
cupies the  position  A,  as  before,  but  is  to  rotate  in  the  direction  indicated  by  the  arrow 
2,  then  the  position  occupied  by  the  eccentric  must  be  as  shown  in  dotted  lines  and 


MODEKX  LOCOMOTIVE   CONSTRUCTION. 


87 


with  its  center  at  C.  Now  if  the  engine  is  to  rotate  at  one  time  in  a  given  direction, 
and  at  another  time  in  an  opposite  direction,  or,  in  other  words,  if  the  motion  of  the 
engine  is  at  any  time  to  be  reversed,  then  we  must  have  some  device  by  which  one 
eccentric  can  be  moved  from  its  position  B  to  (7,  or  we  must  have  two  eccentrics  fixed 
on  the  axle  for  each  slide-valve.  The  latter  method,  namely,  the  use  of  two  eccentrics 
for  each  slide-valve,  has  been  adopted  in  locomotive  engines.  At  present,  and  for  the 
sake  of  simplicity,  we  will  continue  the  investigation  of  the  link  and  its  motion  as  used 
in  a  valve  gear  in  which  no  rocker  is  employed. 

Referring  now  to  Fig.  163,  it  will  be  readily  perceived  that  when  the  engine  is  to 
turn  in  the  direction  as  indicated  by  arrow  1,  then  the  eccentric  drawn  in  full  lines, 
and  whose  center  is  at  £,  and  that  alone,  must  move  the  slide-valve ;  and  when  the 
engine  is  to  rotate  in  the  opposite  direction,  as  indicated  by  the  arrow  2,  then  the 
eccentric  drawn  in  dotted  lines  (and  not  the  other  one)  must  move  the  slide-valve. 
From  the  foregoing  we  conclude  that  in  order  to  reverse  the  engine 
-  .4.  we  must  disengage  one  eccentric  and  engage  the  other,  and  for  this 

purpose  the  link,  as  shown  in  Figs.  164  and  165,  is  employed.  In 
Fig.  164  we  see  that  one  end  of  each  eccentric-rod  is  attached  to  the 
link.  In  this  h'nk  a  slot  or  opening  D  D  is  cut  lengthwise  in 
which  a  block  E,  called  the  link-block,  can  freely  but  accurately 
move  from  one  end  to  the  other  end  of  the  link.  The  piece  F  is 


called  the  saddle  and  is  bolted  to  the  link.  To  the  saddle  the  pin  G  is  forged,  and 
is  called  the  link  saddle-pin.  The  shaft  H  is  called  the  lifting-shaft,  or  the  re- 
versing shaft;  and  the  arms  I,  J  are  called  the  lifting-shaft  arms.  A  pin  is  fast- 
ened to  the  end  of  the  lifting-shaft  arm  /.  This  pin  and  the  link  saddle-pin  G  work 
freely  in  a  piece  K,  called  the  link  hanger;  this  link  hanger  is  simply  a  connection 
between  the  link  and  the  lifting  shaft.  To  the  lifting-shaft  arm  J,  one  end  of  the  reach 
rod  is  attached,  as  shown.  The  other  end  of  the  reach  rod  connects  with  the  reversing 
lever,  which  is  placed  in  the  cab  of  the  locomotive.  The  reversing  lever  is  here  repre- 
sented by  its  center  line  only ;  more  of  this  hereafter.  It  will  readily  be  seen  that,  by 
moving  the  reverse  lever  in  the  direction  as  indicated  by  the  arrow  3,  the  link  can 
be  raised  to  any  position  desired,  and  thus  the  motion  of  the  engine  reversed. 

100.  There  are  two  methods  of  applying  the  link.  First,  it  may  be  applied  as  shown 
in  Fig.  164.  In  this  case,  if  we  move  the  reverse  lever,  we  also  move  the  h'nk  and 
not  the  block,  and  thus  set  the  link  to  any  desired  position.  Of  course,  in  moving  the 
link,  the  end  of  the  eceentrie-rods,  which  are  attached  to  it,  are  carried  with  it. 


88  MODERN  LOCOMOTIVE   CONSTRUCTION. 

Links  which  are  moved  by  the  reversing  lever,  so  as  to  reverse  the  motion  of  the  en- 
gine, are  called  "  shifting  links." 

The  second  method  of  applying  the  link  is  illustrated  in  Fig.  165.  Here  the  revers- 
ing lever  moves  the  valve-rod  link  to  which  the  link-block  is  attached,  but  does  not 
move  the  link ;  or,  in  short,  we  may  say  that,  in  order  to  reverse  the  motion  of  the 
engine,  the  link-block  is  shifted  in  the  link.  In  this  case  the  link  is  called  a  "  stationary 
link." 

By  the  term  "stationary"  is  simply  meant  that  the  link  is  suspended  from  a 
stationary  or  fixed  point ;  the  link  itself  is  not  stationary,  because,  when  the  engine  is 
running,  either  one  or  the  other  eccentric,  or  both,  will  act  upon  the  link,  and  thus  keep 
it  continually  on  the  move. 

101.  It  will  readily  be  seen  by  referring  to  Fig.  164  that  when  a  shifting  link  is 
used,  and  the  engine  is  to  rotate  in  the  direction  indicated  by  arrow  1,  then  the  link 
must  occupy  the  position  as  here  shown ;  that  is,  it  must  have  been  moved  downwards, 
and  for  full  gear  the  eccentric-rod  pin  B2  and  the  center  of  the  link-block  E  will  lie  in 
a  horizontal  line.     If  the  engine  is  to  rotate  in  an  opposite  direction,  then  the  link 
for  full  gear  must  be  lifted  up  until  the  center  of  eccentric-rod  pin  G'2  and  the  center 
of  the  link-block  will  be  in  the  same  horizontal  line. 

Now,  referring  to  Fig.  165,  we  see  that  when  the  stationary  link  is  used,  and  the 
engine  is  to  rotate  in  the  direction  as  indicated  by  arrow  1,  the  link-block,  not  the  link, 
must  be  moved  upwards  until  its  center  and  the  center  of  the  eccentric-rod  pin  B.2  lie 
in  a  horizontal  line,  as  here  shown,  and  when  the  engine  is  to  move  in  an  opposite 
direction,  then  the  link-block  must  be  moved  downwards  until  its  center  and  the  center 
of  the  eccentric-rod  pin  C2  are  again  in  a  horizontal  line  for  full  gear. 

102.  It  may  also  be  of  interest  to  the  reader  to  note  some  of  the  differences  in  the 
construction  of  the  shifting  and  that  of  the  stationary  links.    In  the  former  the  curva- 
ture of  the  link  is  towards  the  axle ;  that  is,  the  center  from  which  the  link  has  been 
drawn  is  located  towards  the  axle.     On  the  other  hand,  the  center  from  which  the  sta- 
tionary link  is  drawn  is  located  towards  the  slide-valve.     Again,  notice  that  in  the 
shifting  link  the  eccentric-rods  are  coupled  to  the  concave  side  of  the  link,  and  in  the 
stationary  link  the  eccentric-rods  are  coupled  to  the  convex  side  of  the  link.     It  can 
also  be  shown  that  when  the  latter  link  is  used  the  lead  of  the  slide-valve  will  be  con- 
stant at  whatever  point  of  the  stroke  steam  may  be  cut  off,  but  when  a  shifting  link  is 
used  the  lead  of  the  slide-valve  will  not  be  constant ;  that  is,  the  earlier  that  the  steam 
is  cut  off  the  greater  will  be  the  lead.     This  we  shall  presently  explain.     We  must  also 
note  the  fact  that  when  a  shifting  link  is  employed  in  a  manner  as  here  shown,  the 
angular  advance  of  the  eccentric  is  found  according  to  the  rule  given  in  Arts.  65,  67, 
but  when  the  stationary  link  is  used  the  angular  advance  of  the  eccentrics  will  be  less. 
In  American  locomotives  the  shifting  link  is  mostly  used,  and  the  stationary  link  is 
seldom  found ;  therefore,  hereafter  we  will  generally  confine  our  attention  to  the  inves- 
tigation of  the  shifting  link. 

103.  From  the  foregoing  the  reader  may  be  led  to  believe  that  the  whole  purpose 
of  the  link  is  to  take  one  eccentric  out  of  gear  and  place  the  other  into  gear ;  and,  in- 
deed, the  writer  believes  that  when  the  link  was  first  discovered  no  one  expected  to  use 
it  for  any  other  purpose.     But  soon  afterwards  engineers  became  aware  of  the  fact  that 


MODERX  LOCOUOTirE   CONSTRUCTION.  gg 

the  link  could  be  used  for  cutting  off  steam  in  the  cylinder  at  different  parts  of  the 
stroke,  and  that  on  account  of  its  simplicity  it  was  particularly  well  adapted  in  locomo- 
tive engines  for  this  purpose.  Hence  we  may  say  that  the  purpose  of  the  link  is  two- 
fold :  first,  because  with  it  the  motion  of  the  engine  can  readily  be  reversed ;  second, 
tlie  point  of  cutting  off  steam  in  the  cylinder  can  easily  be  changed.  Thus,  for  instance, 
it'  the  link  is  placed  in  the  position  as  shown  in  Fig.  164  it  will  in  nowise  affect  the  point 
of  cutting  off  steam  in  the  cylinder ;  that  is,  if  the  eccentrics  are  set  to  cut  off  steam 
at  J  of  the  stroke,  and  the  valve  has  the  proper  amount  of  lap,  the  link  will  not  change 
this  point  of  cutting  off.  Again,  when  the  link  is  raised  up  so  that  the  center  of  the 
eccentric-rod  pin  C2  will  be  in  line  with  the  center  of  the  link-block,  then  the  motion 
of  the  engine  will  be  simply  reversed,  and,  as  before,  the  point  of  cut-off  will  not  be 
interfered  with.  If  now,  on  the  other  hand,  the  link  is  raised  a  short  distance  only,  so 
that  the  center  of  the  link-block  will  be  located,  say,  about  midway  between  the  end  of 
tin'  link  and  the  link  saddle  F,  then  the  travel  of  the  valve  will  be  shortened  and  the 
point  of  cutting  off  steam  in  the  cylinder  will  be  changed. 

DEFINITIONS. 

104.  Since  the  cylinders  are  placed  in  front  of  the  locomotive,  it  follows  that  when 
the  engine  is  traveling  ahead,  the  crank  must  turn  in  the  direction  as  indicated  by  the 
arrow  1,  Fig.  164 ;  but  we  have  seen  that  when  the  crank  rotates  in  this  direction  the 
eccentric  B  must  work  the  valve ;  therefore  the  eccentric  B  is  called  the  forward  eccen- 
tric, and  the  rod  connected  with  it  is  called  the  forward  eccentric-rod.     The  eccentric  C 
is  the  backward  eccentric,  and  the  rod  connected  with  it  is  called  the  backward  eccentric- 
rod.    Again,  when  the  link  occupies  the  position  shown  in  Fig.  164,  or  when  it  occupies 
the  other  extreme  position,  that  is,  when  the  link  is  moved  up,  then  the  link  is  said  to 
be  in  full  gear.     When  the  link-block  stands  midway  between  either  one  of  its  extreme 
positions  and  the  center  of  the  saddle-pin,  the  link  is  said  to  be  in  half  gear ;  and  lastly, 
when  the  center  of  the  link-block  is  in  line  with  the  center  of  the  saddle-pin,  the  link 
is  said  to  be  in  mid  gear. 

The  backward  stroke  of  the  piston  is  that  described  from  the  front  end  of  the  cylin- 
der towards  the  crank,  and  the  forward  stroke  is  that  described  from  the  back  end  of 
the  cylinder  towards  the  front. 

For  the  sake  of  brevity,  and  according  to  custom,  we  shall  hereafter  call  the  dis- 
tanee  from  the  center  of  the  eccentric  to  the  center  of  the  eccentric-rod  pin,  the  length 
of  the  eccentric-rod ;  or,  in  other  words,  we  shall  consider  the  eccentric  and  rod  to  be 
one  piece,  and  therefore  the  distance  from  the  center  B  to  B.,,  or  from  C  to  C2,  Figs. 
1(>4  and  16"),  will  be  the  length  of  the  eccentric-rod. 

By  the  term  radius  of  link  is  meant  the  radius  of  the  arc  I)  I),  drawn  through  the 
center  of  the  opening  of  the  link. 

LEAD   AND   ANOULAB   ADVANCI.    INT    CONNECTION   WITH   LINKS. 

105.  In  Art.  102  we  have  stated  that  when  a  stationary  link  is  employed  the  lead 
remains  constant  at  whatever  point  of  the  stroke  the  steam  may  be  cut  off.     The  truth 


90 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


of  this  will  be  evident  by  referring  to  Fig.  166.  Here,  as  the  position  of  the  slide-valve 
indicates,  the  piston  stands  at  the  beginning  of  the  forward  stroke,  and  consequently 
the  center  of  the  crank-pin  will  be  at  u  on  the  center  line  of  motion  L  M.  In  order  to 
enable  us  to  trace  the  action  of  the  mechanism  of  the  valve  gear  as  clearly  as  possible,  we 
have  represented  the  latter  by  its  center  lines.  All  lines  drawn  in  full  represent  the 
position  of  the  different  parts  of  the  valve  gear  when  the  piston  is  at  the  beginning  of  the 
forward  stroke,  and  consequently  correspond  with  the  position  u  of  the  crank  as  shown. 
All  the  dotted  lines  represent  the  position  of  the  mechanism  when  the  piston  is  at  the 


beginning  of  the  backward  stroke,  and  consequently  will  correspond  with  a  position  of 
the  crank  opposite  to  that  of  u.  The  line  d  k  represents  the  center  of  the  valve-rod  link, 
and  the  distance  from  d  to  k  represents  the  length  of  the  valve-rod  link  from  center  to 
center  of  pins.  The  arc  d  I  represents  the  link  arc,  that  is,  an  arc  drawn  through  the 
center  of  opening  in  the  link.  And  here  again,  for  the  sake  of  simplicity,  we  have 
assumed  that  the  center  of  eccentric-rod  pins  are  located  in  the  same  arc  d  I.  In  this 
case  such  an  assumption  will  not  affect  the  correctness  of  our  reasoning.  The  distance 
between  the  points  d  and  /  on  the  arc  d  I  represents  the  distance  between  the  eccentric- 
rod  pins.  The  circumference  of  the  circle  /  b  represents  the  path  of  the  center  of  eccen- 
tric. The  point /in  this  circumference  represents  the  position  of  the  forward  eccentric, 
and  the  point  I  in  the  same  circumference  represents  the  center  of  the  backward  eccen- 
tric. Both  centers /and  b  are  shown  in  the  correct  relative  positions  to  that  of  the 
crank,  when  the  piston  is  at  the  beginning  of  the  forward  stroke.  When  in  this  posi- 
tion the  points /and  b  will  lie  in  a  line  parallel  to  the  line  S  T,  which  is  drawn  perpen- 
dicular to  L  M.  The  point  f,  represents  the  position  of  the  forward  eccentric,  and  b2 
represents  the  position  of  the  backward  eccentric  when  the  piston  is  at  the  beginning  of 
the  backward  stroke.  These  points/,  and  b.2  will  also  lie  in  a  straight  line  parallel  to  the 
line  S  T.  And  since  the  points /and  b  lie  in  the  same  circumference,  and  also  in  a  line 
perpendicular  to  L  M,  it  follows  that  the  points /and  b  are  equally  distant  from  the 
center  line  of  motion  L  M.  The  same  remarks  apply  to  the  points/  and  b.2.  The  full 
lines  /  d  and  b  I  represent  the  center  lines  of  the  eccentric-rods  when  the  piston  is  at  the 
beginning  of  the  forward  stroke,  and  the  dotted  lines  b.,  /.,  and/,  d2  represent  the  center 
lines  of  the  eccentric  rods  when  the  piston  is  at  the  beginning  of  the  backward  stroke. 
The  link  is  suspended  in  such  a  manner  that  when  the  piston  is  at  the  beginning  of 
the  backward  or  forward  stroke,  the  center  line  of  motion  L  Mwill  pass  midway  between 
the  ends  of  the  link,  or,  in  other  words,  the  line  L  M  will  pass  midway  between  the  points 
d  and  I.  In  stationary  links  the  radius  of  the  link  arc  d  I  is  equal  to  the  length  d  k  of  the 
valve-rod  link.  Now,  since  the  centers  /and  b  of  the  eccentrics  lie  in  a  straight  line 
perpendicular  to  L  M,  and  since  the  lines  /  d  and  b  I  are  equal  in  length,  and  also,  since 
the  lines  fd  and  b  I  when  produced  towards  L  would  form  equal  angles  with  the  line 


MODERX   LOCOMOTIVE    CONSTRUCTION.  91 

/,  .17,  it  follows  that  the  point  k  from  which  the  arc  d  I  is  drawn  will  also  lie  in  the  line 
L  .17.  Consequently,  when  the  link-block  is  lowered,  or,  that  is  to  say,  when  the  center 
d  of  the  valve-rod  link  is  moved  towards  J,  the  point  A;  will  remain  stationary,  and 
therefore  the  lead  will  not  be  changed,  no  matter  what  position  the  link-block  may 
occupy  in  the  arc  d  I.  But  when  the  link-block  is  at  d,  the  valve  motion  is  assumed  to 
be  in  full-gear ;  on  the  other  hand,  when  the  link-block  occupies  a  position  on  the  arc 
d  I  anywhere  between  the  points  d  'and  /,  the  valve  motion  is  not  in  full-gear ;  hence 
the  travel  of  the  valve  is  changed,  and  consequently  the  point  of  cut-off  is  also  changed 
without  changing  the  lead  of  the  valve. 

106.  When  a  shifting  link  is  employed  and  no  rocker  used,  then  the  linear  advance 
of  the  valve  and  the  angular  advance  of  the  eccentric,  measured  as  explained  in  Art. 
(>7,  will  be  equal  to  each  other,  and  consequently  the  angular  advance  of  each  eccentric 
must  be  found  as  shown  in  Arts.  65  and  67. 

In  Art.  102  we  have  stated  that  when  a  stationary  link  is  used  the  angular  advance 
of  the  eccentric  will  be  less  than  that  which  is  necessary  when  a  shifting  link  is 
employed.  In  the  first  place,  then,  let  us  consider  why  this  should  be  so ;  and  secondly, 
let  us  establish  a  method  for  finding  this  angular  advance  of  the  eccentrics,  with  sta- 
tionary link. 

We  have  already  seen  that  when  the  piston  stands  at  the  beginning  of  the  forward 
stroke,  one  end  of  the  valve-rod  link  will  be  at  k,  Fig.  166,  and  when  the  piston  is  at  the 
beginning  of  the  backward  stroke,  the  same  end  of  the  valve-rod  link  will  be  at  n.  The 
distance  between  the  points  k  and  n  on  the  line  L  M  must  be  equal  to  twice  the  linear 
advance  of  the  valve.  Again,  since  the  two  lines  d  k  and  d2  n  are  parallel,  it  follows 
that  the  line  d.2  d,  which  is  drawn  parallel  to  L  M,  must  be  equal  to  the  distance  between 
the  points  «  and  k,  or  in  other  words,  the  distance  between  the  point  d  and  d.,  must  be 
equal  to  twice  the  linear  advance  of  the  valve.  Now  notice  that  when  the  piston  stands  at 
the  beginning  of  the  forward  stroke  the  eccentric-rods  /  d  and  b  I  do  not  cross  each  other ; 
on  the  other  hand,  when  the  piston  stands  at  the  opposite  end  of  the  stroke  the  eccen- 
tric-rods do  cross  each  other,  as  shown  by  the  dotted  lines  b.2  12  and  /,  d.2.  Consequently 
the  angle  formed  by  the  line/rf  (when  it  is  produced  towards  L)  and  the  line  L  M  will 
be  less  than  the  angle  formed  by  the  lines  f,  d.,,  and  L  M,  and  therefore,  on  account  of 
the  inequality  of  these  angles  the  distance  between  the  straight  line  that  may  be  drawn 
through  the  points  f,  b.,  and  the  straight  line  drawn  through  the  points /and  b  will  be 
less  than  the  distance  between  the  points  d,2  and  d.  But  the  angle  formed  between  the 
line  S  Tand  a  straight  line  joining  the  points /and  c  is  equal  to  the  angular  advance 
of  the  eccentric.  Or  again,  the  angle  formed  between  the  line/c  and  the  line  bz  c  is 
equal  to  twice  the  angular  advance  of  the  eccentric.  But  we  have  just  seen  that  the 
distance  between  the  points  b,2  and /is  less  than  that  between  the  points  d.2  and  d,  and 
consequently  the  angular  advance  of  tin-  eccentric  will  be  less  than  the  linear  advance 
of  the  valve ;  and  lastly,  since  in  a  valve  gear  in  which  the  shifting  link  is  used,  as 
shown  in  Fig.  1(>4,  tin-  angular  advance  of  the  eccentrics  is  equal  to  the  linear  advance 
of  the  valve,  it  follows  that  when  a  stationary  link  is  used  the  angular  advance  of 
the  eccentrics  will  be  less  than  that  which  is  necessary  when  a  shifting  link  is  em- 
ployed. 


92 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


METHOD   FOR   FINDING   THE   ANGULAK   ADVANCE. 

107.  In  order  to  show  clearly  how  to  find  the  angular  advance  of  the  eccentric  in  a 
case  as  shown  in  Fig.  166,  we  will  take  the  following  example : 

EXAMPLE  35. — Lap  of  valve  is  1  inch,  lead  -fa  of  an  inch,  travel  of  valve  5  inches, 
length  of  eccentric-rods  3  feet,  and  the  distance  between  the  eccentric-rod  pins  in  the 
link  is  12  inches,  throw  of  the  eccentrics  is  5  inches.  Find  the  angular  advance  of  the 
eccentrics  suitable  for  a  stationary  link. 

Let  c  on  the  line  L  M,  Fig.  167,  be  the  center  of  the  driving  axle.  From  c  as  a 
center  and  with  a  radius  equal  to  2J  inches  (that  is,  half  the  throw  of  the  eccentric) 
describe  the  cii-cle/k  The  circumference  of  this  circle  will  represent  "the  path  of  the 


centers  of  the  eccentrics.  Draw  two  lines  a  and  g  parallel  to  the  horizontal  line  of 
motion  L  M  and  each  line  equally  distant  from  L  M.  The  total  distance  between  the 
lines  a  and  g  must  be  equal  to  that  between  the  centers  of  the  eccentric-rod  pins  in  the 
link,  namely  12  inches.  From  the  center  c  and  with  a  radius  of  3  feet  (which  is  equal 
to  the  length  of  the  eccentric-rods)  draw  a  short  arc  intersecting  the  line  a  in  the  point 
m.  Through  the  point  m  and  the  center  c  draw  a  straight  line  m  c,  and  prolong  it  to 
the  circumference/,  &2,  on  this  line,  fi-om  the  center  c,  and  with  a  distance  of  1-fa  inches 
(which  is  equal  to  the  linear  advance  of  the  valve),  lay  off  the  point  It ;  and  also  with 
the  same  distance  (1  -fa  inches)  lay  off  from  the  center  c  on  the  line  c  m  the  point  i. 
Through  the  points  i  and  h  draw  lines  perpendicular  to  the  line  c  m,  intersecting  the 
circumference  of  the  circle  in  the  points/  and  f2.  The  point /will  be  the  center  of  the 
forward  eccentric  when  the  piston  is  at  the  beginning  of  the  forward  stroke,  and  the  point 
/>  will  be  the  center  of  the  same  eccentric  when  the  piston  is  at  the  beginning  of  the  back- 
ward stroke.  From  the  points  /  and  />  as  centers,  and  with  a  radius  equal  to  the  length 
of  the  eccentric-rod,  describe  small  arcs  intersecting  the  line  a  in  the  points  d  and  d.2. 
Through  the  points  d  and /draw  a  straight  line,  then  this  line  df  will  be  the  center 
line  of  the  forward  eccentric-rod  when  the  piston  is  at  the  beginning  of  the  forward  stroke. 
Through  d.2  and  /,  draw  a  straight  line,  then  this  straight  line  d.2  f2  will  be  the  center  of 
the  same  eccentric-rod  when  the  piston  is  at  the  beginning  of  the  backward  stroke. 

The  distance  between  the  points  d  and  d.,  will  be  equal  to  twice  the  linear  advance 
of  the  valve  very  nearly.  We  say  "  nearly,"  because  this  method  of  finding  the  angular 
advance  is  empirical,  and  can  be  accepted  only  as  an  approximate  method.  Yet  in 


MOVER X  LOCOMOTIVE   COXSTRCCTIOy. 


93 


ordinary  cast's  the  difference  between  the  line  d  d2  and  the  linear  advance  is  inappre- 
ciable, and  even  in  extreme  cases,  such  as  represented  in  the  figure  in  which  the  eccen- 
tric-rods are  comparatively  very  short,  the  result  is  very  nearly  correct.  Yet  in  every 
case  tin*  distance  Ix-twcen  the  points  d.2  and  d  found  by  the  foregoing  method  should 
be  compared  with  the  linear  advance  of  the  valve,  and  when  it  is  found  that  the  dis- 
tance lie)  ween  the  points  d  and  d.,  is  greater  than  twice  the  linear  advance,  the  former 
must  be  corrected  by  changing  the  positions  of  the  points/ and  f2.  The  difference  is 
generally  so  small  that  the  correction  necessary  for  the  positions  of  the  points/and/ 
••an  very  readily  be  seen.  In  this  example  the  linear  advance  of  the  valve  is  l/g  inches, 
and  according  to  the  construction  in  Fig.  167  the  angular  advance  of  the  eccentric  is 
! ,',  of  an  inch  measured  on  a  line  drawn  through  the  point  /  perpendicular  to  the  line 
S  T.  The  point  b  represents  the  center  of  the  backward  eccentric  when  the  piston  is  at 
the  beginning  of  the  forward  stroke,  and  the  point  b.2  represents  the  center  of  the  same 
eccentric  when  the  piston  stands  at  the  beginning  of  the  backward  stroke.  The  position 
of  the  point  b  is  found  by  drawing  through  the  point /a  straight  line/i,  parallel  to  the 
line  S  T.  The  point  b  in  which  the  line/ b  intersects  the  circumference  of  the  circle  is 
the  center  of  the  backward  eccentric.  In  a  similar  manner  we  find  b2  by  drawing  a  line 
through/,  parallel  to  S  T. 

LEAD  WITH   SHIFTING   LINKS. 

108.  In  Art.  105  we  have  seen  that  the  lead  of  the  valve  remains  constant  when  a 
stationary  link  is  employed. 

But  now  let  us  examine  the  state  of  affairs  when  a  shifting  link  is  used ;  by  so 
doing  we  will  find  that  the  lead  of  the  valve  increases  when  the  link  is  moved  from 
full-gear  towards  mid-gear. 

Fig.  168  represents  a  valve  gear  with  a  shifting  link,  and  here  again,  for  the  sake 
of  simplicity,  its  mechanism  is  represented  by  center  lines.  Also,  in  order  to  enable 


us  to  trace  quickly  and  clearly  the  effect  of  the  position  of  the  link  on  the  lead  of  the 
slide-valve,  we  have  shown  the  latter  and  its  seat  above  the  line  />  .17  and  parallel  to  it. 
Tin-  distance  between  the  valve  seat  and  the  line  /,  M  is  immaterial ;  it  can  be  placed 
at  any  convenient  height,  without  affecting  the  correctness  of  our  reasoning;  but  the 


94  MODERN  LOCOMOTITE    CONSTRUCTION. 

distance  between  the  line  S  T  and  the  end  of  the  valve  seat  is  important,  and  should  be 
placed  in  a  position  as  will  be  presently  explained. 

In  Fig.  168  the  point  c  represents  the  center  of  the  driving  axle;  the  circumfer- 
ence /  b  represents  the  path  of  the  centers  of  the  eccentrics — -/the  center  of  the  forward 
eccentric  and  &  the  center  of  the  backward  eccentric.  The  arc  d  I  represents  the  link 
arc,  that  is,  an  arc  drawn  through  the  center  of  the  link  opening ;  and  the  arc  e  g 
represents  the  arc  in  which  the  centers  of  the  eccentric-rod  pins  are  located.  In  this 
figure  the  link  motion  is  shown  to  be  in  full-gear.  The  center  of  the  crank-pin  is  at 
M,  and  consequently  the  piston  will  be  at  the  beginning  of  its  forward  stroke. 

When  the  link  motion  is  in  full-gear,  or  in  mid-gear,  or  in  any  intermediate 
position,  the  point  of  intersection  h  of  the  line  L  M  with  the  arc  d  I  will  always 
represent  the  position  of  the  center  of  the  valve-rod  pin;  and  since  the  distance 
between  the  valve-rod  pin  and  the  slide-valve  is  constant,  it  follows  that  if  we  know 
the  position  of  the  former  we  also  know  the  position  of  the  latter.  If,  therefore, 
through  the  point  h,  a  straight  line  h  i  be  drawn  perpendicular  to  L  M,  and  the 
valve  seat  placed  in  a  position  in  which  the  distance  between  the  line  h  i  and  the 
outer  edge  p  of  the  port  will  represent  the  lead  when  the  link  motion  is  in  full-gear, 
then  we  can  easily  determine  the  amount  of  lead  when  the  link  is  set  to  cut  off  at  any 
other  portion  of  the  stroke.  Thus  for  instance :  Let  the  arc  d  I  represent  the  position 
of  the  link  arc  when  the  link  motion  is  in  full-gear  and  the  piston  at  the  beginning  of 
the  forward  stroke,  and  also  assume  that  the  valve  has  i^  of  an  inch  lead  when  the  link 
motion  is  in  this  position.  Draw  a  straight  line  n  o  any  convenient  distance  above 
and  parallel  to  L  M.  Through  /;,  the  point  of  intersection  of  the  line  L  M  with  the 
arc  d  I,  draw  a  straight  line  h  i  perpendicular  to  L  M ;  the  point  of  intersection  of  the 
line  h  i  with  the  line  n  o  will  represent  the  edge  of  the  valve  as  shown. 

From  the  line  h  i  and  on  the  line  n  o  lay  off  a  point  p  -fa  of  an  inch  from  h  i,  then 
this  point  p  will  represent  the  outer  edge  of  the  steam  port,  and  the  distance  between 
the  point  p  and  the  line  h  i  is  the  lead  when  the  link  motion  is  in  full-gear.  Let  us 
now  assume  that  the  link  has  been  moved  into  mid-gear  as  shown  by  the  dotted  lines, 
but  without  disturbing  the  position  of  the  crank  and  that  of  the  eccentric  centers /and 
b.  Through  7»2,  the  point  of  intersection  of  the  line  L  M  with  the  new  position  of  the 
link  arc  d2  12,  draw  a  straight  line  k.2  i.2  perpendicular  to  L  M;  the  distance  between 
this  line  and  the  outer  edge  p  of  the  port  will  represent  the  lead  when  the  link  is  in  mid- 
gear,  and,  as  will  be  seen,  this  lead  is  greater  than  the  lead  when  the  link  is  in  full-gear. 

109.  In  a  similar  manner  it  can  be  shown  that  the  lead  gradually  increases  when 
the  link  is  moved  from  full-gear  towards  mid-gear.  Again,  by  simply  increasing  the 
length  of  the  eccentric-rods  the  difference  between  the  lead  when  the  link  is  in  full- 
gear  and  the  lead  when  the  link  is  in  mid-gear  is  decreased.  Thus,  for  instance,  making 
the  length  of  the  eccentric-rods  equal  to  twice  the  length  as  before,  but  not  changing 
the  position  of  the  crank  and  the  centers /and  I  of  the  eccentrics,  the  link  will  occupy 
the  position  as  shown  at  x ;  then  by  drawing  the  valve  seat  in  the  correct  place,  follow- 
ing the  same  method  of  construction  as  before,  it  will  be  seen  that  the  difference  in  the 
lead  (or  in  other  words  the  distance  between  the  line  h  i  and  h.2  i-2)  when  in  full-gear 
and  the  lead  when  the  link  is  set  in  mid-gear  is  less  than  when  shorter  eccentric-rods 
are  used.  From  this  we  leara  that  the  magnitude  of  the  variable  character  of  the  lead 


Mf)I>Klt\   LOCOMOTIVE   CONSTRUCTION. 


95 


depends  upon  the  length  of  the  eccentric-rods,  and  that  in  practice,  where  it  is  generally 
desirable  to  keep  the  lead  as  nearly  constant  as  possible,  we  must  make  the  eccentric- 
rods  as  long  as  the  design  of  the  engine  will  admit.  In  locomotives  when  in  full-gear 
the  lead  is  generally  ^  of  an  inch,  sometimes  a  little  less,  and  this  lead  is  increased  to 
£  or  §  of  an  inch,  and  sometimes  even  more,  by  moving  the  link  into  mid-gear. 

In  order  to  avoid  hereafter  any  misunderstanding,  we  again  call  attention  to  the  fact 
that  the  foregoing  remarks  refer  to  link  motions  in  which  rockers  are  not  employed. 

CONNECTION   OP  ECCENTKIC-KODS  TO  THE  LINK. 

110.  It  is  always  desirable  that  locomotive  slide-valves  should  have  some  lead,  no 
matter  in  what  position  the  link  is  placed,  and  it  certainly  would  be  injurious  if  the 
slide-valve  lapped  over  the  steam  port  at  the  beginning  of  a  stroke  of  the  piston.  Now 


to  avoid  having  lap  at  the  beginning  of  a  stroke,  the  eccentric-rods  must  be  correctly 
connected  to  the  link.  Notice,  for  instance,  the  manner  in  which  the  eccentric-rods  are 
connected  to  the  link  in  Fig.  168.  There,  it  will  be  seen,  the  eccentric-rods  do  not  cross 
each  other  when  the  piston  is  at  the  beginning  of  the  forward  stroke.  But  now  let  us 
examine  Fig.  169,  which  represents  precisely  the  same  valve  gear  as  that  shown  in  Fig. 
168,  but  with  this  difference — the  eccentric-rods  are  crossed  when  the  piston  is  at  the 
beginning  of  the  forward  stroke.  Note  the  result.  If  it  is  the  intention  to  use  the  link 
simply  for  the  purpose  of  reversing  the  motion  of  the  engine,  then  this  manner  of  con- 
necting the  eccentric-rods  to  the  link  would  work  very  well ;  but  as  soon  as  the  link  is 
used  for  the  purpose  of  changing  the  point  of  cut-off,  then  this  arrangement  of  eccentric- 
rods  will  have  an  injurious  effect,  particularly  so  in  locomotive  engines,  as  we  can  readily 
see  by  inspecting  Fig.  169.  In  this  figure  the  full  lines  represent  the  mechanism  in 
full-gear,  and  the  dotted  lines  represent  the  same  in  mid-gear.  Drawing  a  line  h  i 
perpendicular  to  the  line  L  M,  and  placing  the  valve  seat  in  its  proper  position  as  ex- 
plained in  connection  with  Fig.  168,  the  distance  between  the  lino  h  i  and  the  outer 
edge  of  the  steam  port,  Fig.  169,  will  be  the  amount  of  lead  when  the  link  is  placed 
in  full-gear.  On  the  other  hand,  when  the  link  is  placed  in  mid-gear,  as  shown  by  the 
dotted  linos,  and  drawing  a  line  h.,  i,  through  the  point  //.,  perpendicular  to  the  line 
L  717,  we  find  that  instead  of  having  lead — as  we  should  have — the  slide-valve  laps  over 
the  steam  port  when  the  piston  is  at  beginning  of  the  stroke,  which  is  a  bad  feature, 
and  must  be  avoided  in  locomotive  construction. 


96  MOltKRX   LOCOMOTII'K    COXliTRl'CTION. 

PKACTICAL   APPLICATION   OF   THE   PRINCIPLES   RELATING   TO   THE  VALVE   MOTION. 

111.  In  our  previous  articles  we  have  endeavored  to  explain  the  mode  of  pro- 
cedure in  laying  out  on  paper  a  simple  valve  gear,  so  that  its  mechanism  can  be 
correctly  proportioned,  drawn,  and  made  in  the  shop. 

After  the  different  parts  of  the  valve  gear  are  finished,  they  must  then  be  correctly 
set  in  the  engine.  Although  the  methods  employed  in  the  shop  for  finding  the  position 
of  the  eccentric  and  other  mechanism  of  the  valve  gear  may  appear  to  be  different  from 
those  employed  for  finding  the  positions  of  the  same  pieces  on  paper,  the  principles  on 
which  these  methods  are  based  do  not  differ. 


In  setting  a  simple  valve  gear  such  as  is  illustrated  in  Fig.  170,  the  great  aim  is  to 
obtain  equal  leads  of  the  slide-valve ;  to  obtain  these  we  must  first  determine  the  cor- 
rect length  of  the  eccentric-rod ;  second,  we  must  find  the  locations  of  the  dead  centers 
of  the  crank ;  and  lastly,  the  correct  positions  of  the  eccentric  on  the  shaft. 

In  order  to  show  clearly  the  practical  method  employed  in  setting  the  valve  gear 
in  an  engine  of  this  kind,  we  will  take  the  following  example : 

EXAMPLE  36. — The  distance,  as  shown  in  Fig.  170,  between  the  center  of  the  crank- 
shaft and  the  line  drawn  midway  between  the  steam  ports  of  the  cylinder  is  8  feet  4 
inches,  the  length  of  the  valve-rod  from  center  of  valve  to  center  of  valve-rod  pin  is  20 
inches,  lap  §  of  an  inch,  lead  -r6  of  an  inch,  and  travel  of  the  valve  5  inches.  In  this 
example  it  must  be  understood  that  the  cylinder,  shaft,  and  the  other  parts  of  the 
engine  have  been  correctly  set  in  line,  and  that  all  we  have  to  do  is  to  set  the 
valve  gear. 

LENGTH   OF  ECCENTRIC-ROD. 

Our  first  duty  is  to  find  the  length  of  the  eccentric-rod ;  on  paper  this  can  be  easily 
accomplished.  Here  we  have  only  to  draw  the  valve  and  rod  in  the  center  of  its  travel, 
and  measure  the  distance  from  the  center  E  of  the  valve-rod  pin  in  mid  position  to  the 
center  of  the  shaft,  which  is  equal  to  100  —  20  =  80  inches,  and  this  distance  of  80  inches 
is  the  length  of  the  eccentric-rod. 

Now  if  the  workmanship  of  all  the  other  parts  of  the  engine  is  positively  per- 
fect, so  that  all  the  dimensions  of  the  mechanism  are  absolutely  correct,  all  that  we 
need  to  do  is  to  make  the  eccentric-rod  from  the  center  of  the  eccentric-strap  to  the 


MODKKX  LOCOMOTIl'K   COXSTRCCTIOX.  97 

editor  of  ihe  valve-rod  pin  SO  inches  long.  But  such  perfect  workmanship  is  seldom 
procured,  ami  therefore  the  eccenti'ic-rods  are  generally  made  in  two  pieces,  namely, 
the  eccentric-strap  and  the  rod  proper,  and  constructed  so  that  the  distance  between 
the  center  of  strap  and  the  center  of  pin  E  can  be  adjusted  to  suit  the  other  parts  of 
the  machinery,  and  thus  enabling  all  to  work  harmoniously  and  correctly. 

To  obtain  in  an  engine  by  measurement  the  distance  from  the  center  of  shaft  to 
the  center  I]  is  often  a  difficult  matter  if  not  impracticable,  and  therefore  the  following 
practical  method  for  finding  the  correct  length  of  the  eccentric-rod  is  employed. 

Fasten  the  eccentric  on  the  shaft  in  a  position  which  will  allow  it  to  be  connected 
to  the  valve-rod.  In  fastening  the  eccentric  in  this  position,  no  attention  need  or 
should  be  paid  to  the  position  of  the  crank.  Place  and  connect  in  position  the  eccen- 
tric-rod, which  we  will  assume  to  be  somewhat  short.  Turn  the  crank-shaft  in  a 
direction  in  which  the  shaft  is  designed  to  run,  and  when  the  valve  arrives  in  the 
position  marked  1,  drawn  in  full  lines  (Fig.  171),  representing  it  to  be  at  one  extreme 
end  of  its  travel,  draw  along  the  edge  a  of  the  valve  a  h'ne  on  the  valve  seat;  again 
turn  the  shaft  in  the  same  direction  as  before,  and  when  the  valve  arrives  in  the 
position  marked  2,  shown  in  dotted  lines,  representing  it  to  be  at  the  other  extreme 
end  of  its  travel,  draw  along  the  edge  b  of  the  valve  a  line  on  the  valve  seat.  Also  on 
this  surface  draw  a  short  line  d  midway  between  the  lines  a  and  b  and  parallel  to  the  same. 
The  distance  from  the  line  d  to  the  line  e  drawn  midway  between  the  steam  ports  indi- 
cates  that  the  length  of  the  eccentric-rod  is  just  that  much  too  short  and  must  be 
increased  by  an  amount  equal  to  this  distance.  If  this  measures  i  of  an  inch  the 
length  of  rod  must  be  increased  by  i  of  an  inch.  Again,  if  the  point  d  had  fallen  on  the 
other  side  of  the  center  line  e,  then  the  distance  between  d  and  e  would  have  indicated 
that  the  eccentric-rod  is  just  that  much  too  long,  and  must  be  shortened  by  an  amount 
equal  to  this  distance. 

In  Art.  61  we  find  that  when  a  valve  has  no  lap  the  center  of  the  eccentric  is 
placed  in  a  line  perpendicular  to  the  center  line  of  motion,  and  in  Art.  67  we  find  that 
when  a  valve  has  lap  the  angular  advance  of  the  eccentric  must  be  laid  off  from  this 
same  line.  Therefore  in  Fig.  171  the  angular  advance  of  the  eccentric  must  be  laid  off 
from  the  line  0  P  drawn  perpendicular  to  the  center  line  of  motion  L  M.  Conse- 
quently, on  paper,  the  position  of  the  eccentric  is  easily  found,  for  we  have  only  to  draw 
a  circle  whose  diameter  is  equal  to  the  travel  of  the  valve,  namely,  5  inches;  and  draw 
a  straight  line/w  parallel  to  0  P  and  {-£  °f  an  inch  (which  is  equal  to  the  lap  and  lead) 
away  from  it.  The  point  /in  which  the  lino/wi  intersects  the  circle  is  the  center  of 
the  eccentric  when  the  crank  is  at  L,  the  crank-shaft  rotating  in  the  direction  as  indi- 
cated by  the  arrow.  But  in  setting  the  valve  gear  in  an  engine,  lines  like  L  M  and 
0  P,  from  and  on  which  measurements  can  be  taken,  would  be  a  difficult  matter  to 
locate,  and  therefore  we  must  seek  another  method,  but  not  new  principles,  for  laying 
off  the  angular  advance  of  the  eccentric.  To  do  so  we  must  find  the  dead  centers  of 
the  crank. 

TO   FIND   THE   DEAD   CENTERS   OF  A   CRANK. 

The  dead  centers  A  and  fi  of  the  crunk-pin,  in  Ki.ii.  17<»,  are  represented  by  the 
points  in  which  the  center  line  of  motion  L  M  intersects  the  circumference  of  the  circle 


98  MODERN  LOCOMOTITE   CONSTRUCTION, 

representing  the  path  of  the  center  of  the  crank-pin.  But,  as  stated  before,  we  cannot 
locate  in  the  engine  the  line  L  M;  we  must,  therefore,  in  this  case  also  adopt  a  practical 
method  by  which  these  dead  centers  can  be  readily  and  correctly  found.  For  the  sake 
of  simplicity  in  our  illustration  we  have  represented  a  crank-disk  instead  of  a  locomo- 
tive wheel.  This  will  not  affect  the  correctness  of  our  reasoning,  for  what  is  true  in 
one  instance  will  also  be  true  in  the  other. 

On  the  crank-disk  describe  arcs  I  c  and  e  d ;  if  the  periphery  of  the  disk  is  turned, 
then  the  arcs  I  c  and  e  d  can  be  described  with  the  aid  of  a  gauge ;  if  the  periphery  is 
not  turned,  then  these  arcs  should  be  described  with  the  aid  of  a  scriber  or  sharp-pointed 
instrument  held  against  the  face  of  the  disk  while  the  shaft  is  revolving  in  its  bearings ; 
but  whichever  way  the  arcs  are  described  these  must  be  true.  Next  turn  the  shaft  in 
the  direction  of  the  arrow,  until  the  crosshead  is  within  a  short  distance  from  the  end 
of  the  stroke,  say  £  of  an  inch ;  while  in  this  position  mark  on  the  slides  a  line  g  even 
with  the  end  k  of  the  crosshead.  Also,  while  the  shaft  and  the  crosshead  is  in  this 
position,  place  a  center-punch  mark  j  on  the  frame  or  any  other  fixed  surface.  From 
this  point  j  as  a  center,  and  with  a  tram  of  any  convenient  length,  j  h,  as  a  radius, 
describe  a  short  arc  intersecting  the  arc  &  c  in  the  point  c  (this  point  c  will  at  this 
instant  coincide  with  the  end  h  of  the  tram,  and  not  as  shown  in  the  figure).  Now  turn 
the  shaft  in  the  same  direction  as  before,  causing  the  crosshead  to  complete  its  full 
stroke  and  part  of  the  return  stroke,  and  when  during  this  motion  the  edge  k  of  the 
crosshead  touches  the  line  g  on  the  slide,  stop  turning  the  shaft,  and  while  in  this  posi- 
tion describe  from  the  point  j  as  a  center,  and  with  the  same  tram  as  before,  a  short  arc 
intersecting  the  arc  I  c  in  the  point  It.  Find  the  point  li  on  the  arc  b  c,  midway  between 
the  points  I  and  c.  Now  turning  the  shaft  into  a  position  in  which  the  ends  of  tram 
will  touch  the  points  h  and  j,  the  crank  will  then  be  on  one  of  its  dead  centers,  as  shown. 

In  a  similar  manner  we  can  find  the  point  i,  but  for  this  purpose  we  must  draw 
another  line,  /,  on  the  slide ;  this  time  I  must  be  drawn  even  with  the  edge  n  of  the 
crosshead  when  it  is  about  one-half  of  an  inch  from  the  beginning  of  the  backward 
stroke.  Then  from  the  point  j,  and  with  the  same  tram-,  and  in  the  same  manner  as 
before,  the  points  d  and  e  are  found,  and  the  point  i,  midway  between  these  points  on 
the  arc  d  e,  is  established.  Turning  the  shaft  into  a  position  in  which  the  points  of  the 
tram  will  touch  the  points  j  and  «,  the  crank  will  then  be  on  the  other  dead  center. 

PRACTICAL    METHOD   OF   FINDING   THE   ANGULAR  ADVANCE   OF   THE   ECCENTRIC. 

When  now  the  crank  is  placed  on  a  dead  center,  the  valve  must  then  be  in  a 
position  in  which  the  steam-port  opening  is  equal  to  the  lead.  Therefore  place  the 
crank  on  a  dead  center,  say  on  A,  and  move  the  eccentric  (which  is  now  assumed  to  be 
connected  to  the  valve-rod  E)  into  a  position  so  that  the  valve  will  have  ^  of  an  meh 
lead,  thereby  giving  the  eccentric  the  correct  angular  advance.  Fasten  the  eccentric 
in  this  position.  If  no  inaccuracies  exist  in  the  valve  gear,  then  by  turning  the  crank- 
shaft we  will  find  the  same  amount  of  lead  when  the  crank  is  at  B.  If  the  valve  has 
not  the  same  lead  at  each  end  of  the  stroke,  then  inaccuracies  do  exist,  which  must  be 
found  and  rectified.  In  setting  the  valve  extreme  accuracy  is  necessary ;  without  this 
failure  will  be  the  result. 

112.  A  direct-acting  valve  gear,  as  shown  in  Fig.  170,  is  not  used  on  locomotives. 


MODEKN  LOCOMOTITK   COXSTRUCTIOy.  99 

Tliis  kind  of  valve  has  been  shown  here  to  enable  us  to  point  out  some  fundamental 
principles  which  must  be  remembered  in  laying  out  any  kind  of  valve  gear.  We  will 
repeat  here  the  most  important  ones  in  an  order  in  which  they  will  present  themselves 
in  laying  out  a  valve  gear  for  any  locomotive : 

1.  The  position  of  the  eccentrics  must  be  laid  off  from  a  line  drawn  perpendicular 
to  the  center  line  of  motion  of  the  valve  gear,  and  not  to  the  center  line  of  crank.  If 
the  center  line  of  motion  coincides  with  the  center  line  of  crank,  then  the  line  drawn 
perpendicular  to  the  former  will  also  be  perpendicular  to  the  latter ;  but  this  is  merely 
a  case  of  coincidence,  and  does  not  prove  that  the  line  from  which  the  positions  of  the 
eccentrics  are  laid  must  be  drawn  perpendicular  to  the  center  line  of  crank.  (See  Art. 
63,  page  4.°,.) 

'2.  When  no  rocker  is  used  the  linear  advance  will  be  equal  to  the  angular  advance 
of  the  eccentric,  the  latter  being  measured  on  a  line  drawn  from  the  center  of  the 
eccentric  perpendicular  to  the  line  from  which  it  is  laid  off.  (See  Art.  67,  page  46.) 

3.  When  no  rocker  is  employed  the  eccentric  will  travel  ahead  of  the  crank.     (See 
Art.  67,  page  46.) 

4.  The  use  of  a  shifting  link  does  not  change  the  angular  advance  of  the  eccentric. 
The  use  of  a  stationary  link  will  change  the  angular  advance  of  the  eccentric.     (See 
Art.  102,  page  88.) 

5.  When  no  rockers  and  links  are  employed  the  throw  of  the  eccentric  will  be 
equal  to  the  travel  of  the  valve.     (See  Art.  55,  page  37.)     When  rockers  with  arms  of 
unequal  lengths  are  used  the  throw  of  the  eccentric  will  not  be  equal  to  travel  of  the 
valve.     When  the  rocker-arms  are  of  equal  length  the  throw  will  be  equal  to  the  travel 
of  the  valve.     (See  Art.  97,  page  82.) 

6.  The  lead  varies  with  a  shifting  link ;  the  lead  is  constant  with  a  stationary  link. 
(See  Art.  102,  page  88.) 

7.  The  eccentric-rods  must  be  connected  correctly  to  a  shifting  link,  otherwise 
there  will  be  no  lead  when  the  link  is  moved  towards  mid-gear.     (See  Art.  108,  page  93.) 

CLASSIFICATION   OF  LINKS. 

113.  When  links  are  classified  with  reference  to  the  manner  of  their  suspension, 
we  have,  according  to  Art.  100,  the  shifting  link  and  the  stationary  link. 

When  these  same  links  are  classified  with  reference  to  their  form,  we  have  the 
following  two  classes,  namely,  the  box  link  as  shown  in  Fig.  172,  and  the  open  link  as 
shown  in  Figs.  173  and  175.  In  American  locomotives  the  former  is  seldom  employed, 
the  open  link  being  the  favorite ;  and  therefore  we  will  consider  the  latter  only. 

The  open  link  can'  again  be  divided  into  two  classes,  namely,  the  solid  link,  as 
shown  in  Fig.  173,  and  the  built-up  link,  generally  called  the  skeleton  link,  shown  in 
Fig.  175.  The  term  "  skeleton  link  "  we  shall  hereafter  adopt  for  this  class  of  links. 

DEFINITIONS. 

In  all  links  the  link  arc  is  an  arc,  as  a  b  c,  drawn  through  the  center  of  the  opening, 
as  shown  in  Fig.  173. 

Length  of  link  is  the  length  of  the  opening  measured  on  a  straight  line  joining 
the  ends  a  and  c  of  the  arc  a  b  c. 


100 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


Radius  of  link  is  the  radius  with  which  the  link  arc  a  b  c  has  been  described,  as 
stated  in  Art.  104. 

Eccentric-rod  pin  arc  is  an  arc,  as  e  d,  drawn  through  the  centers  of  the  eccentric- 
rod  pin-holes  F  F ;  this  arc  is  described  from  the  same  center  as  that  used  in  describing 
the  link  arc  a  b  c. 


!O 


o 


Fig.  172 


11 


'XI 


1C 


Fig.  174 


Fig.  173 


Fig.  175 


The  manner  of  suspending  the  link  and  attaching  the  same  to  the  rocker  is  plainly 
shown  in  Fig.  174.  E  is  the  lower  rocker-arm ;  A,  the  link-block  pin ;  7?,  the  link-block ; 
L,  the  link ;  S,  the  link-saddle ;  P,  the  link-saddle  pin ;  H,  the  link  hanger ;  6',  the  end 
of  the  lifting  shaft  arm ;  and  D,  the  lifting  shaft  pin. 


LINK-BLOCK  PIN. 

114.  The  link-block  pin  A  (Fig.  174)  is  made  of  wrought-iron  case  hardened,  and 
is  fastened  to  the  lower  rocker-arm.  Its  end,  which  fits  into  the  lower  rocker-arm, 
should  be  tapered  and  accurately  fitted  into  the  latter.  The  taper  should  be  the  same 
as  that  of  the  valve-rod  pin,  as  given  in  Art.  90. 

The  diameter  of  the  link-block  pin  at  A  is  generally  made  equal  to  that  of  the 
valve-rod  pin.  Hence,  the  diameters  of  these  pins  will  be : 

For  10  and  11  cylinders If  diam. 

"      12,  13,  and  14    "        li      " 

"     15  and  16  '       "        If      " 

"     17     "     18          "        13      " 

"     19     "     20         •"  li     " 


J/0/*A7f.V  LOCOMOTITE   CONSTRUCTION. 


101 


Comparing  these  figures  with  the  diameters  of  the  pins  in  actual  practice,  it  will  be 
found  that  the  diameters  of  the  pins  given  for  the  smaller  cylinders  agree  very  closely 
with  those  in  use,  and  the  diameters  given  for  17,  18,  19,  and  20  cylinders  are  rather 
large.  But  it  must  be  remembered  that  the  diameters  here  given  have  been  calculated 
for  cylinders  having  steam  ports  suitable  for  piston  speed  of  600  to  800  feet  per  minute, 
which  will  be  required  for  fast  passenger  service,  and  with  a  steam  pressure  of  120 
pounds  in  the  cylinder.  For  freight  engines  in  which  the  steam  ports  are  often 
smaller  than  those  adopted  for  fast  passenger  service,  and  consequently  have  smaller 
slide-valves,  the  diameter  of  these  pins  may  be  somewhat  reduced,  because  they  will 
have  less  work  to  do. 

LINK-BLOCKS. 

115.  The  link-block  is  made  of  wrought-iron  and  is  case  hardened ;  it  works  freely 
and  accurately  on  the  pin  A  (Fig.  174).  For  skeleton  links  the  link-block  is  generally 
made  in  one  piece ;  but  when  the  solid  link  is  used  the  link-block  consists  of  two  or 
three  pieces.  Fig.  176  represents  a  side  view;  Fig.  177  an  end  view;  and  Fig.  178  a 


d  b 


Fly.  177 


fig.  17S 


L 


Fig.  186    a 


'j.  181 


a        a 
Fig.  184 


Fig.  186 


section  of  a  link-block  which  is  made  in  three  pieces,  namely,  the  plates  b  and  d,  and 
the  block  c ;  after  the  block  c  has  been  placed  in  the  opening  of  the  link,  the  plates  b 
and  d  and  the  block  c  are  riveted  together  with  four  g"  rivets  when  link-block  is  large, 
and  with  four  \"  rivets  when  the  link-block  is  small;  the  position  of  the  rivets  is 
shown  in  Fig.  176. 

The  advantage  claimed  for  a  link-block  made  in  this  manner  is,  that  the  curved 
surfaces  a  n  of  the  block  c  and  the  edges  of  the  plates  \>  and  r/ can  lie  finished  in  a 
slotting  machine,  which  in  some  shops  is  more  convenient  to  do  than  turning  the 
curved  surfaces  of  the  block.  Some  master-mechanics  prefer  to  make  the  link-block  in 
two  pieces,  us  shown  in  Figs.  17i),  180,  and  181.  Link-blocks  made  in  this  manner 


102  MODERN  LOCOMOTIVE   CONSTRUCTION. 

consist  of  the  plate  b  and  the  block  c  with  the  projections  or  flanges  d  d  forged  on  to 
it.  Close  to  the  flanges  d  d  small  grooves  e  e,  about  -fa  of  an  inch  deep  and  ^2-  of  an 
inch  wide,  are  turned  into  the  curved  surfaces ;  with  these  grooves  the  surfaces  can  be 
finished  completely  with  an  emery  wheel,  without  using  special  tools  for  finishing  the 
corners  after  the  link-block  has  been  case  hardened.  Sometimes  the  plate  b  is  riveted 
to  the  block,  as  shown  in  Fig.  176,  and  at  other  times  the  plate  I  (Fig.  181)  is  not 
fastened  to  the  block  c  in  any  manner.  In  this  case  the  plate  b,  when  the  link-block  is 
in  position,  must  be  next  to  rocker-arm,  and  is  prevented  from  turning  around  by  a 
pin  f,  f  of  an  inch  diameter,  as  shown  in  Figs.  179,  180 ;  the  block  and  plate  being  held 
together  by  the  link-block  pin.  All  link-blocks,  no  matter  whether  they  are  made  in 
one  piece  or  several  pieces,  are  counterbored,  as  shown  in  the  figures  to  receive  the  head 
of  the  link-block  pin. 

The  length  of  the  block  c  is  generally  from  one  and  one-quarter  to  one  and  one- 
half  times  the  throw  of  the  eccentric.  This  distance  is  measured  on  a  straight  line  join- 
ing the  ends  c  c  of  an  arc  drawn  through  the  center  of  the  block,  as  shown  in  Fig.  176. 

The  thickness  at  g  or  k,  Fig.  176,  between  the  pin  and  the  link  is  generally  -fa  of 
an  inch. 

In  a  number  of  engines  the  plates  b  and  d  extend  beyond  the  ends  of  the  block  c, 
as  shown  in  Fig.  177 ;  this  is  done  to  gain  larger  wearing  surfaces.  But  since  the 
extension  of  these  plates  will  occasionally  cause  trouble  in  oiling  the  link-block,  the 
plates  are  sometimes  cut  flush  with  the  ends  of  the  block,  as  shown  in  Fig.  179. 

The  depth  of  the  flanges  at  h  /*,  Fig.  176,  is  generally  -fa  of  an  inch,  and  we  have 
seen  them  f  of  an  inch  deep ;  but  in  the  latter  case  the  distance  between  the  link  arc 
and  the  eccentric-rod  pin  arc  was  greater  than  desirable.  The  plates  b  and  d  are 
generally  made  §  of  an  inch  thick. 

The  oil  hole  at  i,  half  way  through  the  metal,  is  J  of  an  inch  in  diameter,  and  then 
increased  to  1  or  l£  inches  in  diameter,  to  hold  the  waste  and  oil. 

PROPOKilONS   OF  LINKS. 

116.  With  a  correctly  designed  shifting  link  motion,  we  obtain  an  equal  lead 
when  the  link  is  in  full-gear,  and  very  nearly  an  equal  lead  when  the  link  is  in  half- 
gear;  we  also  obtain  an  equal  cut-off  when  the  link  is  in  half -gear,  and  as  little 
slip  of  the  link  on  the  block  as  possible.  The  slip  is  greater  when  the  link  is  in  full- 
gear  than  when  it  is  in  mid-gear,  and  generally  the  slip  in  forward  gear  will  exceed 
the  slip  in  the  backward  gear.  But  since  there  will  be  always  more  or  less  slip, 
and  since  this  will  cause  wear  and  create  "  lost  motion,"  the  link  must  be  made  of  such 
metal  as  will  enable  it  to  run  as  long  a  time  as  possible  without  wearing  to  any 
appreciable  extent,  thereby  preserving  the  delicacy  of  its  action. 

We  therefore  find  the  majority  of  the  locomotive  links  made  of  wrought-iron, 
case  hardened,  which  gives  a  smooth  and  excellent  service  to  resist  wear. 

During  late  years  cast-iron  links  and  links  cast  of  steel  have  been  adopted  and  used. 
These  will  wear  faster  than  wrought-iron  links  case  hardened.  But  the  lost  motion  caused 
by  wear  is  not  undesirable,  therefore  cast-iron  or  cast-steel  is  mostly  used  for  the  skel- 
eton links.  In  these  links  very  thin  copper  strips  are  inserted  at  (/  (/  y  ,</,  Figs.  184, 185, 


MODERN  LOCOMOTIVE  CONSTRUCTION.  103 

so  that,  when  the  lost  motion  affects  the  correct  action  of  the  link  to  an  extent  which  is 
hurtful  to  the  engine,  a  strip  or  liner  is  taken  out,  and  the  delicate  action  of  the  link 
restored.  Again,  it'  the  wear  of  these  links  becomes  excessive,  they  can  be  easily  replaced 
by  new  links,  as  the  cost  of  these  is  comparatively  small.  Figs.  185  and  186  represent 
the  form  of  a  skeleton  link  made  of  cast-iron,  and  used  on  a  number  of  mogul  engines 
having  cylinders  18  inches  in  diameter.  The  same  form  is  also  adopted  when  the  link 
is  to  be  cast  of  steel.  We  have  met  with  a  few  locomotives  having  cast-iron  links  of  a 
form  similar  to  that  of  a  solid  wrought-iron  link,  such  as  is  shown  in  Fig.  182. 

But  skeleton  links  are  not  always  made  of  cast-iron  or  cast  of  steel ;  often  we  find 
skeleton  links  made  of  wrought-iron  case  hardened,  as  shown  in  Fig.  175,  and  these  are 
preferred  on  many  railroads. 

In  skeleton  links  a  difficulty  is  experienced  in  putting  back  the  end  bolts  after 
some  of  the  liners  have  been  taken  out,  because  these  bolt  holes — which  incline 
towards  each  other — will  then  not  be  in  line ;  therefore,  in  order  to  avoid  this  diffi- 
culty, some  master-mechanics  make  the  form  of  the  links  as  shown  in  Fig.  184,  in 
which  the  bolts  /*  h  are  parallel  to  each  other. 

117.  The  eccentric-rod  pins  will,  in  a  comparatively  short  time,  wear  the  holes  of  the 
link  oblong ;  and  therefore,  in  order  to  preserve  the  link  as  long  as  possible,  the  holes 
e  and  f,  Fig.  182,  are  bushed ;  the  bushing  is  made  of  wrought-iron  about  ^  or  |  of 
an  inch  thick,  case  hardened,  and  then  forced  into  the  link,  usually  with  a  hydraulic 
pressure  of  four  tons.     This  bushing  is  used  in  wrought-iron,  and  also  in  cast  links. 
When  the  wear  of  the  pin  and  bushing  becomes  so  great  as  to  affect  the  action  of  the 
link,  the  pin  and  bushing  can  be  easily  and  cheaply  replaced  by  new  ones.     Sometimes 
we  find  links  in  which  the  bushing  has  been  fitted  loosely  in  the  holes ;  in  these  cases 
the  bushing  is  slightly  longer  than  the  width  of  the  link,  and  held  fast  in  the  eccentric- 
rod  jaw  by  tightening  the  nut  of  the  eccentric-rod  pin,  allowing  the  bushing  to  move 
freely  in  the  link.     Loose  bushing  is  better  adapted  for  cast  links  than  for  wrought- 
irou  case-hardened  links,  because  the  former  can  be  more  readily  and  in  less  time 
rebored,  or  replaced  by  a  new  link  if  necessary,  and  with  less  expense  than  the  latter. 

118.  The  eccentric-rod  pins  are  also  case  hardened.     The  cross-sectional  area  of 
one  of  these  pins  should  not  be  less  than  half  the  area  of  the  rocker-pins  given  in  Art. 
114 ;  locomotive  builders  generally  make  the  area  of  an  eccentric-rod  pin  a  little  larger 
than  half  the  area  of  rocker-pin,  so  as  to  obtain  a  larger  bearing  surface.     The  diameter 
of  the  eccentric-rod  pins  for  engines  having  cylinders  19  or  20  inches  in  diameter  is 
usually  1J  inches;  for  engines  having  cylinders  16,  17,  or  18  inches  in  diameter,  l£; 
and  for  smaller  engines,  1  inch. 

119.  For  locomotives  having  cylinders  10  inches  in  diameter  and  upwards,  the 
distance  between  the  centers  of  the  eccentric-rod  pins  e  undf,  Fig.  182,  generally  varies 
from  9  to  10  inches ;  and  for  locomotives  having  cylinders  16  inches  in  diameter  and 
upwards,  the  distance  between  these  eccentric-rod  pin  centers  generally  varies  from  11 
to  12  inches;  sometimes,  but  rarely, this  distance  is  13  inches.     The  distances  between 
the  eccentric-rod  pin  centers  shoiild  not  be  made  less  than  those  here  given,  because 
if  we  do  so  the  slip  of  the  link  on  the  block  will  lie  increased.     Neither  can  these  dis- 
tances be  made  much  longer,  because  generally  the  room  under  the  locomotive  will  not 
admit  longer  links. 


104 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


120.  In  locomotives  having  cylinders  10  inches  in  diameter  and  upwards,  the  throw 
of  the  eccentric  is  from  3£  to  4  inches ;  and  in  locomotives  having  cylinders  16  inches  in 
diameter  and  upwards,  the  throw  of  the  eccentric  is  from  4.J  to  5£  inches,  oftener  5 
inches.    Now,  comparing  the  throw  of  the  eccentrics  with  the  distances  between  the 
eccentric-rod  pin  centers,  we  find  that  this  distance  varies  from  2%  to  2£  times  the 
throw  of  the  eccentric.     Hence,  in  designing  a  locomotive  link  we  may  make  the  dis- 
tance between  the  eccentric-rod  pin  centers  equal  to  2J  or  2j  times  the  throw  of  the 
eccentric.     Although  this  is  an  empirical  rule,  it  is  a  good  rule  to  adopt,  provided  it 
does  not  make  the  link  too  long. 

121.  The  distance  between  the  eccentric-rod  pin  arc  and  the  link  arc  must  not 
be  greater  than  necessary;    it  should  be  such  as  will  allow  -fa  of  an  inch  clear- 
ance between  the  flanges  of  the  link-block  and  the  ends  of  the  eccentric-rods.     By 

increasing  this  distance  we  also 
increase  the  slip,  which  must 
be  avoided.  In  ordinary  loco- 
motive practice  this  distance 
varies  from  2J  to  3  inches,  and 
occasionally  reaches  3£  inches. 
122.  The  length  of  the  link, 
that  is,  the  distance  from  c  to  d, 
Fig.  182,  should  be  sufficiently 
great  to  allow  the  center  of  the 
link-block  to  be  placed  in  line 
with  the  center  of  either  one  of 
the  eccentric-rod  pins,  leaving  a 
clearance  sufficient  for  the  slip, 
so  that  when  running  in  this 
gear  the  link-block  will  be  pre- 
vented from  coming  in  contact 
of  an  inch  for  the  least  amount  clear- 


Fly.  183 


Fig.  182 


with  the  end  of  the  link  opening.     In  fact, 

ance  between  the  link-block  and  end  of  link  opening  is  preferable. 

Consequently,  to  determine  the  length  of  a  link,  we  must  know  the  distance 
between  eccentric-rod  pins,  the  length  of  the  link-block,  the  maximum  slip,  and  the 
desired  amount  of  clearances.  The  sum  of  these  items  will  be  the  length  of  the  link. 

EXAMPLE  37. — The  distance  between  the  eccentric-rod  pins  is  12  inches ;  the  length 
of  the  link-block  is  G  inches ;  the  maximum  slip  is  1  £  inches ;  and  the  desired  clearance 
at  either  end  must  not  be  less  than  J  of  an  inch. 

12"  +  6"  +  Ijf"  +  4"  +  4"  =  19&"  =  length  of  link. 

The  length  of  links  in  the  different  locomotives  having  cylinders  16  inches  in 
diameter  and  upwards  varies  from  18  inches  to  19 J  inches,  rarely  exceeding  the  latter 
dimension. 

123.  The  radius  of  links  in  nearly  all  locomotives  is  equal  to  the  distance  between 
the  center  of  the  main  driving  axle  and  the  center  of  link-block  pin  (sometimes  called 
the  lower  rocker-arm  pin)  when  the  latter  stands  ill  the  center  of  its  travel. 


MODEKX   LOCOMOTIVE    CONSTHUCTION.  105 

[t  has  been  found  that,  with  this  radius,  the  variation  of  the  lead  is  sensibly  equal 
for  the  front  and  bark  strokes  of  the  piston.  Sometimes,  when  greater  accuracy  in 
the  equalization  of  the  lead  is  required,  this  radius  of  the  link  is  made  a  little  shorter. 

1:24.  In  order  to  obtain  the  breadth  B,  Fig.  183,  and  the  thickness  T,  Fig.  182,  of 
a  wrought-iron  link,  we  should  know  the  pressure  of  the  valve  against  its  seat ;  but 
since  the  existing  data  is  not  sufficient  to  determine  this  pressure  accurately,  we  will 
assume  that  the  friction  of  the  valve  on  its  seat — and  which  the  link  has  to  overcome 
in  moving  the  valve — is  proportional  to  the  total  steam  pressure  on  the  back  of  the 
valve.* 

Consequently,  for  the  purpose  of  obtaining  these  dimensions  of  the  link  we  will 
adopt  tlif  same  rule  as  that  used  for  finding  the  principal  dimensions  of  an  eccentric, 
given  in  Art.  98. 

Therefore,  for  finding  the  breadth  B  and  thickness  T  of  a  wrought-iron  link  we 
use  the  same  units  found  in  Art.  98,  and  multiply  the  unit  by  the  numbers  given  in 
the  following  table : 

TABLE  13. 

Breadth  B  of  wrought-iron  link  =  unit  x  1.62 
Thickness  T          "          "       "     =  unit  x     .81 

EXAMPLE  38. — Find  the  breadth  B  and  the  thickness  T  of  a  wrought-iron  link  suit- 
able for  a  consolidation  engine  having  cylinders  20  inches  in  diameter ;  the  length  of 
valve  is  10  inches ;  breadth  of  the  same  20  inches ;  and  pressure  of  the  steam  in  the 
steam-chest  120  pounds  per  square  inch. 

We  find  in  Art.  98  that  the  unit  for  this  size  of  slide-valve  is  1.54,  hence : 

1.54  x  1.62  =  2.49"  =  breadth  of  link. 
1.54  x     .81  =  1.24"  =  thickness  T. 

EXAMPLE  39. — Find  the  breadth  B  and  the  thickness  T  of  &  wrought-iron  link 
suitable  for  an  eight-wheeled  passenger  locomotive  having  cylinders  10  inches  in  diam- 
eter ;  slide-valve  being  6  inches  long  and  llj  inches  wide ;  steam  pressure  in  steam- 
chest  120  pounds  per  square  inch. 

We  find  in  Art.  98  that  the  unit  for  this  size  of  slide-valve  is  .91,  hence : 

.91  x  l.(52  =  1.47"  =  breadth  of  link. 
.91  x     .81  =    .73"  =  thickness  T. 

Now,  comparing  the  dimensions  obtained  in  Example  38  with  the  dimensions  of 
the  link  in  Figs.  182  and  183,  which  is  a  drawing  of  a  link  used  in  a  consolidation  engine 
with  cylinder  20  inches  in  diameter,  lately  built  and  now  in  active  service,  we  find 
that  these  dimensions  agree  very  closely.  It  will  also  be  found  that  the  dimensions  of 
other  wrought-iron  links  obtained  by  this  rule,  suitable  for  locomotives  having  cylinders 
13  inches  in  diameter,  and  others  having  cylinders  of  larger  diameter,  up  to  20  indies, 
will  agree  very  closely  with  the  dimensions  of  the  links  in  locomotives  of  the  foregoing 
sixes  at  present  in  active  service. 

*  It  should  be  understood  that  tin-  total  steam  pressure  on  the  back  of  the  valve  is  greater  than  the  pressure  of 
the  valve  agaiust  its  Beat  (see  Art.  82). 


106 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


But  the  dimensions  of  links  for  smaller  locomotives  obtained  by  our  rule  are  less 
than  the  dimensions  of  links  made  according  to  the  present  practice  of  locomotive 
builders. 

Take,  for  instance,  Example  39.  In  this  we  find  that  the  breadth  of  the  link  suit- 
able for  locomotives  having  cylinders  10  inches  in  diameter  is  1.47  inches,  say  1  \  inches, 
whereas  the  breadth  of  links  in  locomotives  of  this  size,  and  at  present  in  active 
service,  is  If  and  sometimes  2  inches.  But  when  it  is  necessary  to  build  a  very  light 

locomotive,  the  writer  believes  that  links  proportioned  by 
this  rule  will  give  good  satisfaction,  although  they  will  wear 
somewhat  faster  than  those  having  a  greater  width. 

125.  The  tendency  in  modern  locomotive  construction  is 
to  make  the  saddle-pin  longer  than  formerly.  This,  in  the 
writer's  opinion,  is  a  great  improvement.  The  saddle-pin 
for  locomotives  having  cylinders  17  up  to  20  inches  in 
diameter  is  now  generally  6  inches  long,  as  shown  in  Fig. 
183.  For  smaller  locomotives  the  length  of  saddle-pin  is 
decreased;  a  saddle-pin  4  inches  long  will  work  very  sat- 
isfactorily in  locomotives  with  cylinders  10  inches  in  diam- 
eter. These  saddle-pins  should  not  be  made  shorter,  unless 
compelled  to  do  so  by  the  narrow  gauge  of  the  road. 
The  diameter  of  the  saddle-pin  is  usually  about  one-eighth  of  an  inch  less  than  the 
diameter  of  the  link-block  pin. 

Figs.  187  and  188  represent  a  link  hanger.  The  dimensions  given  are  suitable  for 
locomotives  having  cylinders  19  inches  and  others  having  cylinders  20  inches  in 
diameter.  Usually  the  holes  are  bushed  with  wrought-iron  ferrules  case  hardened. 


Fig.  187 


LIFTING  SHAFT. 

126.  Fig.  189  represents  an  end  view  and  Fig.  190  a  plan  of  the  "  lifting  shaft," 
or  sometimes  called  the  "  reverse  shaft."  Two  arms  B  B — one  for  each  link — are 
forged  to  the  shaft  D ;  the  holes  for  the  lifting-shaft  pins  A  A  in  the  end  of  these 
arms  are  tapered;  the  taper  should  be  the  same  as  that  of  the  valve-rod  pin  (see 
Art.  90). 

The  case-hardened  lifting-shaft  pins  A  A  are  made  to  fit  these  holes  very  accu- 
rately. On  these  pins  the  link  hangers  vibrate.  The  arm  E  is  generally  forged  to  the 
shaft  D ;  occasionally  it  is  keyed  to  the  shaft.  The  hole  at  C  for  the  reach-rod  pin  is 
bushed  with  a  wrought-iron  ferrule,  case  hardened,  usually  from  £  to  YS  of  an  inch 
thick.  The  reach-rod  pin,  which  connects  the  reach-rod  to  the  arm  E,  is  straight  and 
case  hardened.  The  other  end  of  the  reach-rod  is  connected  to  the  reversing  lever  at 
(7,  as  shown  in  Fig.  192.  The  line  L  B  (Fig.  192)  represents  the  center  line  of  the 
reverse  lever  when  it  stands  in  full-gear  forward,  and  the  line  A  B  represents  the 
center  line  of  the  reverse  lever  when  it  stands  in  full-gear  backward ;  and  when  the 
reverse  lever  stands  in  center  of  the  arc  as  shown,  the  link  motion  is  said  to  be  in 
mid-gear.  Now  when  the  reverse  lever  (which  is  connected  by  the  reach-rod  to  the 
lifting  shaft)  is  moved  from  A  to  L  the  motion  of  the  engine  will  be  reversed ;  or  if 


MODEKX  LOCOMOTIVE   CONSTRUCTION. 


107 


Reach  Hod 

--/ax 

X. 


Fig.  189 


the  reverse  lever  is  moved  to  any  intermediate  position,  the  travel  of  the  valve  will 
!»•  reduced,  and  steam  in  the  cylinder  will  be  cut  off  sooner. 

The  diameter  of  the  lifting  shaft  (Figs.  189  and  190)  and  the  size  of  its  arms  must 
be  sufficiently  large  to  prevent  the  shaft  and  arms  from  springing.  The  dimensions 
given  in  the  figures  are  those  of  lifting  shafts  generally  used  in  locomotives  which  have 
cylinders  20  inches  in  diameter.  For  smaller  locomotives,  which  have  cylinders  10 
inches  in  diameter,  the  lifting-shaft  arms  B  B  measured  close  to  the  shaft  are  usually 

2 £  inches  wide  and  f  of  an 
inch  thick ;  the  arm  E  is 
2£  inches  wide  and  £  of  an 
inch  thick ;  and  the  shaft 
2  inches  in  diameter. 
These  dimensions  are 
gradually  enlarged  as  the 
diameter  of  the  cylinder  is  increased. 

The  location  of  the  lifting  shaft  and  the 
length  of  its  arms  B  B  will  influence  the  equal- 
ization of  the  cut-off ;  therefore  it  is  very  im- 
portant to  assign  the  correct  position  to  the 
lifting  shaft  and  make  the  arms  B  B  of  the 
proper  length.  How  to  find  the  position  of  the 
lifting  shaft  and  the  correct  length  of  the  arms 
B  B  will  be  explained  later.  The  length  of 
the  arm  E  is  generally  limited  by  the  design 
of  the  locomotive;  that  is,  the  length  of  this 
arm  must  be  such  as  will  prevent  the  reach- 
rod  from  coming  in  contact  with  other  parts  of 
the  engine. 

The  center  lines  of  the  arms  B  B,  and  that 
of  the  arm  E,  do  not  often  stand  at  right  an- 
gles to  each  other  as  shown ;  they  should  have 
the  following  relative  positions:  The  center 
line  A  F  of  the  arm  E  should  stand  perpen- 
dicular to  the  reach-rod  when  the  link  motion 
is  in  mid-gear ;  and  the  center  lines  of  the  arms  B  B  should  then  stand  in  the  center 
of  their  total  vibration.  This  will  allow  the  end  of  the  arm  E  to  pass  through 
equal  arcs  on  each  side  of  the  line  A  F  during  the  time  the  links  are  moved  from 
full-gear  forward  to  full-gear  backward.  The  short  arm  F  is  usually  forged  to  the 
shaft  I)\  occasionally  it  is  bolted  to  the  shaft.  To  this  arm  F  a  spring  counter- 
balance is  attached  which  acts  against  the  weight  of  links,  hangers,  etc.,  relieving  the 
engineer  of  considerable  hard  work  in  reversing  the  engine,  and  enabling  him  to  move 
the  reverse  lever  as  easily  in  one  direction  as  in  the  other.  Sometimes,  for  the  purpose 
of  counterbalancing  the  weight  of  the  links,  hangers,  etc.,  volute  springs  are  used,  as 
shown  in  Fig.  191,  but  the  writer  believes  that  a  half  elliptic  spring,  as  shown  in  Figs. 
189  and  190,  will  give  better  satisfaction. 


108 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


127.  Volute  springs,  as  shown  in  Fig.  191,  are  made  of  steel  3  inches  wide,  ^  of 
an  inch  thick,  the  springs  are  6  to  6£  inches  long  for  large  locomotives.  For  small  loco- 
motives these  springs  are  made  of  steel  3  inches  wide,  J  of  an  inch  thick,  and  the  same 
length  as  that  of  the  larger  ones.  The  cast-iron  casing  around  these  springs  is 
generally  bolted  to  the  yoke  brace. 

Elliptic  springs  are  usually  made  nearly  as  long  as  the  space  between  the  frames 
will  allow.  For  large  engines,  4'  8£"  gauge,  the  length  of  these  springs  is  usually 

40  inches  before  compres- 
sion, and  having  5  or  G 
leaves  of  steel  2£  inches 
wide  and  -fg  thick. 

When  these  half  ellip- 
tic springs  are  used,  the 
rod  G  (Fig.  189)  should 
be  attached  to  the  arm  F 
in  a  manner  as  shown, 
which  will  allow  the  spring 
to  be  tightened  or  loosened 
without  disconnecting  the 
rod  G  from  the  spring. 

Reach-rods  for  large 
locomotives  are  usually 
made  of  2|"  x  f"  iron; 
and  for  smaller  locomo- 
tives 2"  x  f  "  iron. 

REVERSE   LEVER. 

128.  The  design  of  an 
engine  and  the  position  of 
its  driving  wheels  deter- 
mines the  location  of  the 
reverse  lever.  Generally, 
in  engines  having  a  foot- 
plate the  lower  end  of  the 
lever  can  be  attached  to 
the  same ;  in  consolidation 
engines  or  hard-coal  burn- 
ers we  are  generally  com- 
pelled to  attach  the  lever 
to  the  frame.  In  all  cases 
the  reverse  lever  is  located  on  the  right-hand  side  of  the  engine.  The  reverse  lever  is 
usually  made  of  wrought-iron ;  but  when  the  part  of  the  lever  below  the  arc  D,  Fig.  193, 
is  very  crooked — which  often  occurs — then  in  order  to  save  labor  in  forging,  the  lower 
part  is  sometimes  made  of  cast-iron  and  bolted  to  the  upper  part,  which  is  made  of 
wrought-h'on.  The  form  of  the  lower  end  of  the  reverse  lever  is  determined  also  by 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


109 


the  design  of  the  engine.  In  consolidation  engines  it  often  happens  that  the  reverse 
lever  has  to  move  between  one  of  the  rear  driving  wheels  and  the  boiler,  and  there- 
fore its  lower  part  has  to  be  made  comparatively  thin  and  very  wide,  as  shown  in  Fig. 
192.  In  engines  which  have  the  reverse  lever  attached  to  the  foot-board,  the  lower 
cin I  of  the  lever  is  shaped  as  shown  in  Fig.  193,  and  its  thickness  is  the  same  through- 
out ;  usually  5  of  an  inch  for  large  engines,  and  §  of  an  inch  for  smaller  ones. 

The  total  length  of  the  reverse  lever  and  the  location  of  the  reach-rod  pin  in  the 
same  must  be  such  as  will  allow  the  top  of  the  reverse-lever  handle  0,  Fig.  192,  to 
move  through  a  distance  of  about  4  feet  in  large  engines;  and  through  a  distance  of 
about  3  feet  6  inches  in  smaller  engines  during  the  time  that  the  link  is  moved  from 
full-gear  forward  to  full-gear  backward.  Now,  since  the  distance  through  which  the 
link  is  moved  is  equal  to  the  distance  between  the  centers  of  eccentric-rod  pins,  and 
since  these  are  placed  from  11  to  12  inches  apart  in  large  engines,  and  from  9  to  10 
inches  in  small  engines  (Art.  119),  we  may  say  that  the  distance  through  which  the 
reverse-lever  handle  moves  should  be  about  four  times  the  distance  through -which  the 
link  moves. 

The  diameter  of  the  reverse-lever  pin  7?,  Fig.  192,  is  usually  1  inch  for  small 
locomotives  and  li  inches  for  large  ones.  These  pins  are  case  hardened. 

The  hole  C  for  the  reach-rod  pin  in  the  reverse  lever  is  sometimes  bushed  with  a 
case-hardened  bushing. 

129.  The  arcs  D  are  usually  made  of  steel,  fastened  to  the  boiler,  or  to  the  foot- 
plate and  running  board.     In  a  number  of  engines  two  arcs  are  employed,  one  on  each 
side  of  the  lever,  as  shown  in  Fig.  192.     In  other  locomotives  one  arc 

only  is  used,  which  passes  through  an  opening  in  the  reverse  lever  as 
shown  in  Fig.  194.  When  a  single  arc  is  used  it  is  made  comparatively 
wide,  as  will  be  seen  by  comparing  Fig.  192  with  Fig.  194.  Whether 
to  use  two  arcs  or  the  single  arc  is  a  matter  of  choice  and  judgment. 
The  arcs  should  be  placed  as  high  as  the  design  of  engine  will  permit ; 
by  so  doing  more  notches,  F  F  (Fig.  192),  can  be  cut  in  the  arcs,  with 
sufficient  metal  for  strength  between  them,  than  can  be  cut  in  arcs 
placed  lower  down.  The  notches  F  F  receive  the  latch  G.  This  latch 
is  connected  to  the  latch-handle  // by  the  links  7,  so  that,  when  the 
latch-handle  is  pressed  towards  the  handle  of  the  reverse  lever,  the 
latch  (i  will  be  lifted  out  of  the  notch,  and  when  the  pressure  on  the 
latch-handle  ceases,  the  spring  K  presses  the  latch  into  the  notch.  With 
this  arrangement  the  lever  can  be  placed  and  held  in  any  desired 
position.  The  writer  believes  that  it  will  give  better  satisfaction  by 
placing  the  latch  G  (which  slides  in  the  clamp  M)  in  front  of  the  re- 
verse lever,  as  shown  in  Fig.  192,  and  not  in  the  rear  of  the  reverse 
lever,  as  shown  in  Fig.  193.  By  adopting  the  former  method  the  reverse 
lever  will  press  against  the  latch;  but  by  placing  the  latch  in  the  rear 
of  the  lever,  the  tendency  will  be  to  pull  the  reverse  lever  away  from  the  latch,  which 
will  in  a  short  time  cause  the  lever  to  rattle  and  interfere  with  the  correct  action  of 
the  link  motion. 

130.  Master-mechanics  differ  in  opinion  in  regard  to  the  number  and  the  position 


Fig.  1U4 


UNIVERSITY  OF  CALIFORNIA 
DEPARTMENT  OF  CIVIL  ENGINEERING 


110 


MODERN   LOCOMOTIVE    CONSTRUCTION. 


of  the  notches  in  the  arcs.  First :  A  number  of  master-mechanics  prefer  the  notches 
arranged  in  a  manner  which  will  hold  the  reverse  lever  in  positions  that  cause  the 
steam  to  be  cut  off  in  the  cylinder  at  some  full  number  of  inches  of  the  stroke.  Con- 
sequently, we  find  arcs  with  notches  cut  in  such  positions  as  will  cause  the  steam  to 
be  cut  off  at  6,  9, 12, 15, 18,  and  21  inches  of  the  stroke ;  or  at  6, 8,  10, 12, 15, 18,  and  21 
inches  of  the  stroke.  Besides  these  notches  one  notch  is  cut  in  the  arcs  to  hold  the 
link  in  mid-gear.  With  this  arrangement  of  notches  a  difficulty  arises,  namely,  it  is 

often  found  that  when  a  particular  notch — 
say  the  6-iuch  notch — holds  the  reverse  lever, 
the    cylinders    do    not    receive    a   sufficient 
amount  of  steam  to  haul  the  train,  and  when 
the  reverse  lever  is  moved  to  the  next  notch 
— the  9-inch  notch — the  cylinders  receive  too 
much  steam,  and  therefore  the  steam  has  to 
be  throttled,  causing  the  locomotive  to  work 
under  disadvantages.     To  overcome  this  dif- 
ficulty, May's  Reverse  Lever  Latch  has  been 
invented,  by  which  a  finer  gradation  is  ob- 
tained without  changing  the  notches.     This 
latch  is  shown  in  Figs.  195  and  196,  and,  as 
will  be  seen,  is  a  very  simple  device.     The 
only  difference  between  this 
and  the  ordinary  latch  is, 
that  the  former  is  a  double 
latch    instead    of    a   single 
one  ;    consequently,    it   can 
easily  be  applied  to  the  re- 
verse levers  at   present   in 
use  without  any  change  in 
the    levers   or  arcs.     Now, 
since  a  finer  gradation  is  not 
a  matter  of  convenience,  but 
it  is  a  saving  of  fuel,  the  advantages  of  May's 
latch,  or  some  equally  good  device,  will  easily 
be  perceived. 

But  while  some  master-mechanics  will 
insist  on  having  the  notches  cut  in  the  fore- 
going manner,  others  believe  that  whether 
steam  is  cut  off  at  full  inches,  or  a  fractional 
number  of  inches  of  the  stroke,  is  of  no  consequence ;  hence  these  master-mechanics 
will  cut  as  many  notches  in  the  arc  as  there  is  room  for,  and  as  close  together  as  the 
strength  of  metal  will  allow.  Fig.  193  shows  an  arc  with  notches  cut  in  this  manner. 
The  distance  between  the  centers  of  these  notches  is  half  an  inch;  sometimes  the 
notches  are  cut  closer  than  this.  With  notches  cut  in  this  manner  a  very  fine  grada- 
tion of  cut-off  is  obtained,  and  fuel  saved. 


310DEKX   LOCOMOTIVE    CONSTRUCTION. 


Ill 


VALVE  GEARS  WITH  ROCKERS. 

131.  Heretofore  we  have  shown,  theoretically  and  practically,  how  to  set  the 
(•reentries  in  simple  valve  gears  in  which  rockers  are  not  employed,  and  in  which 
the  connections  between  the  eccentric  and  valve  are  direct,  and  also  in  which 
the  center  line  of  motion  of  the  valve  gear  coincided  with  that  of  the  piston. 

Let  us  now  continue  the  subject  of  setting 
the  eccentrics,  in  the  following  order:  First,  how 
to  find  the  position  of  an  eccentric  in  a  valve 
gear  in  which  a  rocker  with  arms  of  equal  length 
is  used,  and  in  which  the  center  line  of  motion  of 
the  valve  gear  coincides  with  the  center  line  of 
motion  of  the  piston.  Second,  how  to  find  the 
position  of  the  eccentric  in  a  valve  gear  in  which 
a  rocker  whose  arms  are  not  of  equal  length  is 
used,  and  in  which  the  center  line  of  motion  of 
the  valve  gear  coincides  with  that  of  the  piston. 
Third,  to  find  the  position  of  an  eccentric  in  a 
valve  gear  in  which  a  rocker  is  used,  and  in  which 
the  center  line  of  motion  of  the  valve  gear  does 
not  coincide  with  that  of  the  piston. 

TO  FIND  THE  POSITION  OF  AN  ECCENTRIC  IN  A  VALVE 
GEAR  HAVING  A  ROCKER  WHOSE  ARMS  ARE  OF  EQUAL 
LENGTH,  AND  THE  CENTER  LINE  OF  MOTION  OF  THE 
VALVE  GEAR  COINCIDING  WITH  THAT  OF  THE  PISTON. 

In  order  to  make  this  subject  as  plain  as 
possible,  let  us  take  the  fol- 
lowing example : 

EXAMPLE  39a. — Lap  of 
valve,  ii  of  an  inch;  lead,  Vs 
of  an  inch;  travel  of  valve,  5 
inches ;  length  of  each  rocker- 
arm,  10  inches ;  find  the  posi- 
tion of  the  eccentric. 

In  Art.  67  we  have  explained  how  to  find  the 
position  of  an  eccentric  in  a  simple  valve  gear  in 
which  no  rocker  is  employed,  and  in  Art.  61  we 
have  pointed  to  the  fact  that  hi  simple  valve 
gears  of  this  kind  the  eccentric  must  travel  ahead 
of  the  crank. 

It  is  now  to  be  shown  that,  when  a  rocker  is  interposed  between  the  eccentric  and 
valve  without  making  any  other  changes  in  a  simple  valve  gear,  the  eccentric  must 
1'ollow  the  crank,  instead  of  traveling  ahead  of  the  same;  and  it  is  also  to  be  shown 


Fig.  195 


Fig.  196 


112 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


that,  when  the  rocker  whose  arms  are  equal  in  length  is  used,  the  angular  advance  of 
the  eccentric  will  be  laid  off  in  a  different  direction  from  that  in  a  simple  valve  gear ; 
but  the  amount  of  angular  advance  will  remain  the  same  in  both  cases. 

In  order  to  point  out  clearly  the  reason  why  this  should  be  so,  we  have  illustrated 
in  Fig.  197  two  connections  between  the  valve  and  eccentric.  1st.  The  upper  connec- 
tion, marked  "  Case  1,"  is  that  in  which  no  rocker  is  employed.  2d.  The  lower  connec- 
tion, marked  "  Case  2,"  is  that  in  which  a  rocker  is  interposed. 

In  Case  1  the  center  C  of  the  crank-shaft  is  in  line  with  the  valve ;  in  Case  2  the 
center  C2  of  the  crank-shaft  is  situated  below  the  valve  so  as  to  admit  a  rocker. 

In  Art.  61  we  have  seen  that  the  angular  advance  of  the  eccentric  is  laid  off  from 
a  line  drawn  perpendicular  to  the  center  line  of  crank.  This  method  is  also  applicable 


to  the  example  now  under  consideration ;  but  since  this  will  not  give  correct  results 
in  laying  out  all  valve  gears,  such  as  are  to  be  considered  hereafter,  it  will  be  best  first 
to  establish  a  rule  which  can  be  applied  to  all  cases.  It  is  this : 

RULE  17. — The  angular  advance  of  the  eccentric  must  be  laid  off  from  a  line 
drawn  perpendicular  to  the  line  of  motion  of  the  valve  gear  (see  Art.  63).  Hence,  before 
we  can  comply  with  this  condition  so  as  to  find  the  position  of  the  eccentric  in  Case  1, 
or  in  Case  2,  Fig.  197,  we  must  draw  the  center  line  of  motion  of  the  valve  gear.  To 
do  this  in  Case  1,  we  draw  a  line  L  M  through  the  center  C  of  the  shaft  in  a  direction 
in  which  the  valve  moves.  This  line  L  M  will  be  the  center  line  of  motion  of  the  valve 
in  Case  1,  and  agrees  with  the  definition  given  in  Art.  63.  In  Case  2  we  draw  a  line 
L2  M.,  through  the  center  C.,  of  the  shaft,  and  tangent  to  the  arc  s  t,  described  by  the 


MOIH-:i!.\    I.IH'OMOTH'K   CONSTRUCTION 

center  7?  of  the  lower  rocker-arm  pin.*  This  line  LI  M2  will  be  the  center  line  of 
motion  of  the  valve  gear  in  Case  2. 

Now  let  us  apply  the  method  given  in  Art.  61  for  finding  the  position  of  the  eccen- 
tric in  Case  1,  Fig.  197.  From  the  center  C  of  the  shaft,  and  with  a  radius  of  2£ 
inches  (equal  to  half  the  throw  of  the  eccentric  in  our  example),  describe  a  circle  fdg; 
the  circumference  of  this  circle  will  represent  the  path  of  the  center  of  the  eccentric. 
Through  the  center  C  of  the  shaft  draw  a  line  d  e  perpendicular  to  L  M.  Let  A  repre- 
sent the  position  of  the  center  of  the  crank-pin  when  the  crank  is  on  the  dead  center, 
or,  in  other  words,  when  the  piston  is  at  the  beginning  of  the  stroke. 

Now,  since  the  conditions  in  our  example  demand  that  the  center  line  of  the  valve 
gear  shall  coincide  with  that  of  the  piston,  it  follows  that  the  center  A  of  the  crank- 
pin  must  lie  in  the  line  L  M — and  when  in  this  position  the  valve  must  have  opened 
the  steam  port  ^  of  an  inch,  which  is  equal  to  the  given  amount  of  lead — and  occupy 
the  position  as  shown  in  the  figure.  To  find  the  position  of  the  eccentric  which  will 
correspond  with  that  of  the  valve,  and  enable  the  shaft  to  revolve  in  the  direction 
indicated  by  the  arrow,  we  continue  our  construction  as  follows :  From  the  center  C 
on  the  lino  L  M,  away  from  the  crank-pin  J,  lay  off  a  point  h  ;  the  distance  between 
(J  and  h  must  be  equal  to  the  sum  of  the  lap  and  lead,  namely  1  inch.  Through  the 
point  //  draw  a  straight  line  h  x  parallel  to  the  line  d  e,  cutting  the  circumference  / d  g 
in  the  point  x ;  this  point  x  will  be  the  required  center  of  the  eccentric,  and  will  travel 
ahead  of  the  crank ;  the  valve  will  open  the  steam  port  more  and  more  during  the 
time  the  shaft  revolves  through  a  certain  distance,  and  then  close  the  port,  as  it 
should  do.  In  Art.  61  we  have  drawn  the  line  d  e  perpendicular  to  the  center  line  of 
crank.  In  this  example  we  have  drawn  the  Hue  d  e  perpendicular  to  the  center  line 
of  motion ;  but  since  the  center  line  of  crank  and  the  line  of  motion  coincide,  the 
result  is  the  same. 

The  position  of  the  eccentric  when  a  rocker  is  used,  as  shown  in  Case  2,  Fig.  197, 
is  found  in  the  following  manner :  For  the  sake  of  convenience  and  easy  comparison, 
let  ns  draw  the  center  line  of  motion  L2  M2  of  the  valve  gear  parallel  to  L  M  in  Case  1 ; 
a  No  let  us  place  the  center  G'2  of  the  shaft  in  a  line  drawn  through  C  perpendicu- 
lar to  the  line  L  M,  and  low  enough  to  admit  a  rocker  with  arms  each  10  inches 
long.  Let  o  represent  the  fixed  center  of  the  rocker-shaft,  and  let  the  line  i  k,  drawn 
through  o  perpendicular  to  L.,  M.,,  represent  the  center  line  of  the  rocker-arms  when 
these  stand  midway  of  their  travel,  corresponding  to  the  position  of  the  slide-valve 
when  the  latter  stands  in  a  central  position,  not  indicated  in  these  illustrations. 
Now,  it  must  be  evident  that  when  the  valve  stands  in  the  position  as  shown  in 
the  figure  it  has  moved  1  inch  (the  sum  of  the  lap  and  lead)  out  of  its  central  position, 
and  consequently  the  center  of  the  upper  rocker-arm  pin  must  have  moved  out  of  its 
central  position  the  same  amount  in  a  horizontal  direction  (not  measured  on  the  arc 
described  by  the  center  of  the  pin),  and  therefore  the  center  of  the  pin  will  be  at  /when 
the  valve  stands  in  the  position  as  shown.  Through  the  centers  /  and  o  draw  a  straight 
line  /  R,  cutting  the  arc  s  <,  described  by  the  lower  rocker-pin,  in  the  point  R.  This 

"  Drawing  this  center  line  of  motion  tangent  t.>  the  arc  described  by  the  lower  rocker-arm  pin  is  not  absolutely 
correct,  but  is  near  enough,  and  generally  considered  so.  fur  all  practical  purposes  in  locomotive  cons! met ion,  or  in 
engines  having  eccentric-rods  of  the  ordinary  length ;  that  is,  engines  not  having  very  short  eccentric-rods. 


114  MODERN  LOCOMOTIVE   CONSTRUCTION. 

point  R  will  be  the  position  of  the  center  of  the  lower  rocker-pin  when  the  upper  pin 
is  at  I ;  and,  as  will  be  seen,  these  pins  will  then  be  located  in  the  opposite  sides  along 
the  line  i  k ;  but  the  distance  between  the  center  H  and  the  line  i  k  will  be  the  same  as 
that  between  the  center  I  and  the  line  i  k,  namely,  1  inch,  because  the  rocker-arms  are 
of  equal  lengths.  Also  notice  that  as  the  pin  I  travels  in  the  direction  of  the  arrow  2, 
as  it  should  do,  the  pin  R  will  travel  in  an  opposite  direction,  indicated  by  the  arrow  3. 
When  the  valve  stands  in  the  position  as  shown,  the  crank-pin  A2  in  Case  2  will  be  on 
the  same  side  of  the  shaft  as  A  in  Case  1,  and  will  lie  in  the  line  L2  Mz,  because, 
according  to  the  condition  in  our  example,  the  center  line  of  motion  of  the  valve  gear 
coincides  with  that  of  the  piston.  Therefore  the  following  construction  will  give  us 
the  position  of  the  eccentric  to  correspond  with  that  of  the  crank.  Through  the  center 
C2  of  the  shaft  draw  a  line  d2  e2  perpendicular  to  L2  M2 ;  and  from  the  center  (72,  and 
with  a  radius  of  2^  inches,  describe  a  circle^  d2g2;  the  circumference  of  this  circle 
will  represent  the  path  of  the  center  of  the  eccentric.  From  the  center  C2  on  the  line 
L2  M2,  and  towards  the  crank-pin  A2,  lay  off  a  point  h2 ;  the  distance  between  the 
center  C2  and  the  point  h2  must  be  equal  to  1  inch,  because  the  horizontal  distance 
between  the  point  _R  and  the  line  i  A;  is  1  inch.  Through  the  point  h2  draw  a  line  h2  x2 
parallel  to  the  line  d2  e2,  and  cutting  the  circumference  f2  d2  gz  at  the  point  x2 ;  this 
point  x2  will  be  the  required  position  of  the  eccentric  in  Case  2  when  the  crank-pin  is 
at  A2  and  the  shaft  rotating  in  the  direction  indicated  by  the  arrow.  If  in  Case  2 
we  had  found  the  position  of  the  eccentric  in  precisely  the  same  manner  as  that  em- 
ployed in  Case  1,  and  had  placed  the  eccentric  at  y  and  thus  caused  the  eccentric  to 
travel  ahead  of  the  crank  as  in  Case  1,  a  movement  in  the  wrong  direction  would  have 
been  communicated  to  the  rocker-pin  R,  which  would  make  the  valve  close  the  steam 
port  at  this  particular  time  instead  of  opening  the  same,  as  it  should  do.  Also  notice 
that  y  is  one  end  of  the  diameter  of  the  circle  f2  d2  gz,  and  X2  is  the  other  end  of  the 
same  diameter. 

132.  From  this  we  learn  that  in  a  valve  gear  in  which  a  rocker  with  arms  of  equal 
lengths  is  introduced  the  eccentric  must  be  placed  in  a  position  directly  opposite  to 
that  of  an  eccentric  in  a  valve  gear  in  which  no  rocker  is  used ;  also,  when  the  amount 
of  lap  and  lead  in  Case  1  is  the  same  as  that  in  Case  2,  then  the  angular  advance  in  both 
cases  will  be  equal,  although  laid  off  in  opposite  directions. 

When  two  eccentrics  and  a  link  are  to  be  used,  as  in  locomotives,  then,  in  order  to 
find  the  position  of  the  second  eccentric,  prolong  the  line  h2  x2  so  as  to  cut  the  circum- 
ference^ d2  </2;  this  point  of  intersection  will  be  the  center  of  the  second  eccentric. 
(See  Art.  99.) 

POSITION   OF   ECCENTRICS   WHEN   A   KOCKER   WITH   ARMS   OF   UNEQUAL   LENGTHS   IS   USED. 

133.  In  Fig.  197  we  have  shown  the  position  which  the  eccentrics  must  occupy 
when  the  lengths  of  the  rocker-arms  are  equal. 

If,  however,  the  lower  rocker-arm  is  made  either  longer  or  shorter  than  the  upper 
arm,  then  the  position  of  the  eccentrics  on  the  shaft  must  be  changed  from  that  posi- 
tion they  would  occupy  when  the  arms  of  the  rocker  are  of  equal  lengths. 

EXAMPLE  40. — The  length  of  the  lower  rocker-arm  is  Hi  inches ;  the  length  of  the 


MOVERS    I.OfOMOTIVK   COXSTRVCTIOS. 


115 


upper  arm,  0  inches;  throw  of  eccentric,  5  inches;  lap,  ft  of  an  inch;  lead,  ^  of  an 
inch  ;  the  center  line  of  motion  of  the  valve  gear  coincides  with  that  of  the  piston;  it 
is  required  to  find  the  position  of  the  eccentrics. 

Fig.  19H.  Draw  the  center  line  /  k;  this  line  will  represent  the  center  line  of  the 
rocker-arms  when  these  stand  midway  of  their  travel.  On  the  line  i  k  locate  any  point 
o  to  represent  the  center  of  the  rocker-shaft.  From  the  center  0,  and  with  a  radius 
equal  to  9  inches,  describe  an  arc  u  v  to  represent  the  path  of  the  center  I  of  the  upper 
rocker-pin ;  also  from  the  center  o,  and  with  a  radius  equal  to  11J  inches,  describe  an 
arc  s  t  to  represent  the  path  of  the  center  of  the  lower  rocker-pin.  On  the  arc  u  v  lay 


off  a  point  I ;  the  distance  between  the  line  i  k  and  the  point  I  must  be  equal  to  the  sum 
of  the  lap  and  lead,  namely  1  inch,  measured  on  a  line  perpendicular  to  i  k,  and  not  on 
the  arc  «  r.  Through  the  point  I  and  the  center  o  draw  a  straight  line  /  R,  cutting  the 
arc  s  t  in  the  point  7?.  Draw  L  M,  the  center  line  of  motion  of  the  valve  gear,  perpen- 
ilicular  to  the  line  i  k  and  tangent  to  the  arc  s  t.  On  the  line  L  M  lay  off  the  center  C 
of  the  shaft.  When  the  valve  stands  in  the  position  as  shown  in  the  figure,  the  crank- 
pin  will  be  at  A,  or,  in  other  words,  the  shaft  C  will  be  between  the  crank-pin  and  the 
rocker.  Through  the  center  6' draw  a  straight  line  de  perpendicular  to  L  M;  also 
from  the  center  C,  and  with  a  radius  equal  to  half  the  throw,  namely  2£  inches,  describe 
a  circle ;  the  circumference  of  this  circle  will  represent  the  path  of  the  center  of  eccen- 
trics. From  the  center  C  and  on  the  line  L  M  lay  off  a  point  h ;  the  distance  between 
these  points  must  be  1J  inches ;  through  the  point  h  draw  a  line  parallel  to  the  line  d  e, 
and  cutting  the  circumference  fdg  in  the  points  #  and  y.  The  point  x  will  be  the 
center  of  the  eccentric  when  the  crank-pin  A  has  to  move  in  the  direction  of  the  arrow; 
and  the  point  y  would  bo  the  center  of  the  eccentric  if  the  crank-pin  A  had  to  move 
in  a  direction  opposite  to  that  of  the  arrow.  If  two  eccentrics  and  a  link  are  to  be 
employed,  then  one  eccentric  is  placed  at  x  and  the  other  at  y. 

Since  the  valve  has  -[{!  of  an  inch  lap  and  ,',,  of  an  inch  lead,  the  linear  advance  of 
the  valve  must  be  1  inch;  that  is,  when  the  valve  is  in  the  position  as  shown,  it  has 
traveled  1  inch  away  from  its  central  position;  and  since  the  valve  is  connected  to  the 
upper  rocker-arm,  the  distance  between  the  center  I  and  the  line  i  k  was  made  equal  to 


MODERN  LOCOMOTIVE   CONSTRUCTION. 

1  inch.  According  to  Art.  97,  the  distance  between  the  line  i  k  and  the  center  K  of  the 
lower  rocker-pin  must  be  greater  than  1  inch,  because  the  lower  rocker-arm  is  longer 
than  the  upper  one.  In  our  present  example  the  distance  between  the  line  i  k  and  the 
center  E  is  l£  inches.  But  the  eccentric-rod  is  connected  to  the  lower  rocker-arm,  and 
therefore  the  distance  between  the  center  C  and  the  point  h  must  be  1J  inches,  as  we 
have  made  it.  Hence,  lengthening  the  lower  rocker-arm  necessitated  an  increase  in  the 
angular  advance.  If  the  lengths  of  the  rocker-arm  had  been  equal,  the  distance  between 
C  and  h  would  have  been  1  inch,  or,  in  other  words,  the  centers  x  and  y  of  the  eccentrics 
would  have  been  placed  1  inch  away  from  the  line  d  e,  instead  of  Ij  inches  as  shown 
in  Fig.  198.  In  the  same  manner  it  can  be  shown  that,  when  the  length  of  the  lower 
rocker-arm  is  less  than  the  length  of  the  upper  arm,  then  the  angular  advance  of  the 
eccentric  will  be  less  than  the  linear  advance  of  the  valve. 

From  this  example  we  learn  that,  when  the  lower  rocker-arm  is  longer  than  the 
upper  one,  the  angular  advance  will  be  greater  than  the  linear  advance ;  and  when  the 
lower  rocker-arm  is  shorter  than  the  upper  one,  the  angular  advance  is  less  than  the 
linear  advance — in  short,  the  angular  advance  of  the  eccentric  is  equal  to  the  distance 
between  the  central  position  of  the  lower  rocker-pin  and  that  in  which  it  will  stand 
when  the  piston  is  at  the  beginning  of  the  stroke.* 

POSITION   OF   THE  ECCENTRIC  WHEN   A   ROCKER    IS    USED   AND    THE   CENTER   LINE   OP   MOTION 
OF   THE   VALVE   GEAR  DOES   NOT   COINCIDE   WITH  THAT   OF   THE   PISTON. 

134.  EXAMPLE  41. — The  length  of  each  rocker-arm  is  10  inches ;  lap,  f  f  of  an  inch ; 
lead,  ^g  of  an  inch ;  throw  of  the  eccentric,  5  inches ;  center  line  of  motion  of  the  valve 
gear  does  not  coincide  with  that  of  the  piston  ;  to  find  the  position  of  the  eccentric. 

Fig.  199.  In  this  figure  the  axis  of  the  cylinder  is  assumed  to  be  in  a  line  with  the 
center  C  of  the  shaft ;  that  is,  if  the  axis  of  the  cylinder  is  prolonged  towards  the 
shaft,  it  will  pass  through  the  center  C.  Hence  the  line  N  P  will  be  the  center  line  of 
motion  of  the  piston.  Again,  when  the  crank  is  on  a  dead  center,  the  crank-pin  must 
lie  in  this  line  N  P ;  and  when  the  valve  has  opened  the  steam  port  -fa  of  an  inch,  that 
is  to  say,  when  the  valve  has  -fa  inch  lead,  as  shown  in  the  figure,  the  center  of  the 
crank-pin  must  be  at  A.  Let  o  represent  the  center  of  the  rocker-shaft.  From  the 
center  o,  and  with  a  radius  equal  to  the  length  of  the  lower  rocker-arm,  namely  10 
inches,  describe  the  arc  s  t ;  also  from  the  center  o,  and  with  the  same  radius  as  before, 
describe  the  arc  u  v.  Through  the  center  C  draw  the  line  L  M  tangent  to  the  arc  s  t. 
Then  L  M  will  be  the  center  line  of  motion  of  the  valve  gear ;  and,  as  will  be  seen, 
the  center  line  of  motion  L  M  does  not  coincide  with  the  center  line  of  motion  N  P 

Cases  of  this  kind,  in  which  one  end  of  the  center  line  of  motion  of  the  valve  gear 
is  depressed,  are  not  of  rare  occurrence  in  locomotive  construction ;  we  frequently  have 
to  do  this  in  order  to  give  sufficient  clearance  between  the  lifting-shaft  arms  or  the 
link  and  the  boiler  when  the  valve  gear  is  placed  in  full-gear  back.  But  when  this 

*  It  should  be  remembered  that  increasing'  the  length  of  the  lower  rocker-arm,  and  leaving  the  throw  of  the 
eccentric  the  same,  the  travel  of  the  valve  will  be  decreased.  Also  by  decreasing  the  length  of  the  lower  rocker 
arm  without  changing  the  throw  of  the  eccentric,  the  travel  of  the  valve  will  be  increased.  Therefore  care  anil 
thought  must  be  given  to  the  subject  when  the  lower  rocker-arm  is  made  longer  or  shorter  than  the  upper  rocker-arm 


MtlliKHX    LOCOMOTIVE    COXSTRVCTION. 


117 


Fig.ZOO 


expedient  is  resorted  to,  we  must 
also  make  a  change  in  the  relative 
positions  of  the  rocker-arms  on  the 
shaft,  as  shown  in  Fig.  200.  In  this 
figure  it  will  be  noticed  that  the  cen- 
ter lines  of  the  rocker-arms  do  not 
lie  in  one  straight  line,  as  shown  in  all 
our  previous  figures,  but  that  these 
arms  incline  towards  each  other.  By 
giving  the  rocker-arms  these  posi- 
tions on  tho  shaft  we  will  preserve 
the  identity  and  symmetry  of  their 
motion.  The  relative  positions  of 
the  rocker-arms  are  found  in  the 
following  manner  :  Through  the  cen- 
ter o,  Fig.  199,  draw  a  line  o  k  per- 
pendicular  to  L  M;  this  line  will 
represent  the  center  line  of  the 
lower  rocker-arm  when  it  stands  midway  of  its 
travel.  Also  through  the  center  o  draw  the  line 
o  i  perpendicular  to  the  center  line  of  the  valve-rod  ; 
the  line  o  i  will  represent  the  center  line  of  the  up- 
per rocker-arm  when  it  stands  midway  of  its  travel. 
The  lines  o  i  and  o  k  show  the  required  amount  of 
inclination  of  the  rocker-arms  towards  each  other. 

To  draw  the  rocker  in  a  position  corresponding 
to  that  of  the  valve  at  the  beginning  of  the  stroke, 
lay  off  from  the  line  i  o  on  the  arc  u  ?>,  towards  the 
valve,  a  point  /;  the  distance  between  this  point  / 
and  the  line  i  o  must  be  equal  to  the  linear  advance, 
namely  1  inch.  Through  the  point  /  and  the  center 
ft  draw  a  straight  line  lo;  this  line  will  be  the  center 
line  of  the  upper  rocker-arm  when  in  a  position  cor- 
responding to  that  of  the  slide-valve.  From  the  line 
o  k  on  the  arc  s  t,  and  towards  the  center  C  of  the 
shaft,  lay  off  a  point  E  ;  the  distance  between  the  point 
/.'  and  the  line  o  k  must  also  be  equal  to  the  linear  ad- 
vance (1  inch),  because  the  length  of  the  upper  rocker- 
arm  is  equal  to  that  of  the  lower  one.  The  line  o  R 
represents  the  center  line  of  the  lower  rocker-arm 
when  in  a  position  corresponding  to  that  of  the  valve. 
From  the  center  C,  and  with  a  radius  equal  to  2J 
inches,  describe  the  circle  filff  ;  the  circumference  of 
this  circle  will  represent  the  path  of  the  center  of  ec- 
centric. Through  the  center  ('of  the  shaft  draw  a 


118  MODERN  LOCOMOTIVE   CONSTRUCTION, 

line  d  e  perpendicular  to  the  line  L  M;  from  the  same  center  C  lay  off  on  the  line 
L  M,  towards  the  crank-pin  A  a  point  h;  the  distance  between  the  points  C  and  h 
must  be  equal  to  the  linear  advance  (1  inch).  Through  the  point  h  draw  a  line  x  y 
parallel  to  the  line  d  e,  cutting  the  circumference  /  d  g  at  the  points  x  and  y.  The 
point  x  will  be  the  center  of  the  eccentric  when  the  crank-pin  A  is  to  travel  in  the 
direction  of  the  arrow ;  and  the  point  y  will  be  the  center  of  the  eccentric  when  the 
crank-pin  A  is  to  travel  in  the  direction  opposite  to  that  of  the  arrow.  If  a  link  is  to 
be  used  so  that  the  motion  of  the  crank-shaft  can  be  reversed,  then  the  point  x  will  be 
the  center  of  one  eccentric,  and  the  point  y  the  center  of  the  other  eccentric. 

Now  notice  in  this  case  the  angular  advance  of  the  eccentric  is  laid  off  from  the 
line  d  e,  which  is  not  perpendicular  to  the  center  line  A  C  of  the  crank.  From  this 
we  learn  that  the  angular  advance  must  be  laid  off  from  a  line  di*awn  perpendicular  to 
the  center  of  motion  of  the  valve  gear,  as  stated  in  Rule  17,  and  this  rule  holds  true  in 
all  cases.  On  the  other  hand,  the  expressions,  "  the  eccentric  is  set  at  right  angles  to 
the  crank,"  and  "  the  angular  advance  is  laid  off  from  a  line  drawn  perpendicular  to 
the  crank,"  are  true  only  in  cases  in  which  the  center  line  of  motion  L  M  of  the  valve 
gear  coincides  with  the  center  line  A  C  of  the  crank. 

LAYING   OUT  A   LOCOMOTIVE   VALVE   GEAR. 

135.  In  Fig.  199  we  have  shown  a  valve  gear  without  a  link.     Now,  adding  a  link 
will  not  change  the  position  of  the  eccentrics,  neither  will  it  make  any  difference  in 
the  positions  of  the  rocker-arms ;  the  off-set  in  the  arms  and  the  position  of  the  rocker 
are  not  interfered  with — in  fact,  no  change  whatever  will  be  required  excepting  a  change 
in  the  length  of  the  eccentric-rods.     Hence  all  the  remarks  relating  to  the  valve  gear 
shown  in  Fig.  199  are  true  also  for  a  valve  gear  in  which  a  link  is  used. 

136.  In  regard  to  the  position  of  the  center  o  of  the  rock-shaft  in  any  valve  gear, 
it  may  be  said  that  it  is  usually  located  in  the  most  convenient  position,  which  in 
the  meantime  will  give  as  long  eccentric-rods  as  possible.     Hence  we  find  the  rocker 
placed  either  in  front  of  the  yoke-brace  or  in  the  rear  of  it,  as  shown  in  Fig.  29.    In 
either  case  care  must  be  taken  to  place  the  rocker  far  enough  away  from  the  yoke- 
brace  to  give  sufficient  clearance  between  the  latter  and  the  link  when  hooked  up ; 
again,  it  often  happens  that,  when  we  attempt  in  ten-wheeled  engines  to  place  the 
rocker  in  front  of  the  yoke-brace,  the  link  will  strike  the  engine  truck  frame,  and 
under  these  circumstances  we  are  compelled  to  place  the  rocker  as  shown  in  Fig.  29. 
The  vertical  distance  from  top  of  frame  to  the  center  of  rock-shaft  is  usually  deter- 
mined by  the  location  of  the  valve-rod,  and  when  the  boiler  is  set  compai-atively  low, 
we  may  have  to  lengthen  the  upper  rocker-arms  so  as  to  lower  the  position  of  the 
rocker  for  the  purpose  of  obtaining  sufficient  clearance  between  the  bottom  of  the 
boiler  shell  and  the  top  of  link  and  hanger  when  the  latter  are  placed  in  full  backward 
gear.     The  preceding  remarks  indicate  that  for  determining  the  position  of  rocker-box 
computations  are  not  required,  but  good   judgment  guided  by  experience   must  be 
exercised. 

137.  The  correct  working  of  the  valve  not  only  depends  on  the  correct  position  of 
the  eccentrics,  but  it  will  also  depend,  when  a  link  is  used,  on  the  position  of  the 


v»i>Kn\  T.ocouoTirE  CONSTRUCTION. 

saddle-pin  on  the  link,  the  length  of  the  lifting-shaft  arms  from  which  the  links  are 
suspended,  and  the  position  of  Ilie  lifting  shaft. 

The  position  of  the  driving  axle  in  the  pedestal  will  also  affect  to  a  small  extent 
the  equality  of  the  cut-off,  and  since  this  axle  is  free  to  move  up  and  down  in  the 
pedestal,  the  question  which  presents  Itself  is:  Where  shall  we  place  the  driving 
axle  for  the  purpose  of  laying  out  the  valve  gear?  The  axle  should  be  drawn  in  a 
position  corresponding  to  that  which  it  will  have  when  the  engine  is  in  first-class 
working  order;  and  this  position  can  be  taken  from  Figs.  271  to  279,  in  which  the 
positions  of  axles  in  the  pedestals  for  different  sizes  of  engines  in  working  order  are 
dearly  indicated.  After  the  axle  and  rocker-box  have  been  located,  we  are  then  ready 
for  laying  out  the  valve  gear,  and  in  order  to  show  plainly  the  manner  of  doing  so,  we 
shall  take  the  following  example  and  work  out  the  solution  in  the  same  way  as  many 
(1  raftsmen  will  do;  the  only  difference  being  that  the  draftsman  will  work  out  the 
whole  solution  in  one  diagram,  whereas  we  shall  use  three  to  enable  us  to  point  out  the 
construction  more  clearly. 

EXAMPLE  42. — It  is  required  to  lay  out  a  valve  gear  such  as  is  shown  in  Fig.  29. 
This  gear  is  to  be  used  on  an  eight- wheeled  passenger  engine  with  a  piston  stroke  of  24 
inches.  The  length  of  each  rocker-arm  is  10  inches;  throw  of  eccentrics,  5  inches, 
which  will  make  the  eccentricity  equal  to  2i.  inches ;  lap,  |  inch ;  lead,  /„  inch ;  length 
of  link,  18  inches ;  distance  between  centers  of  eccentric-rod  pins  in  link,  12  inches ; 
length  of  link-hanger,  13  inches;  horizontal  distance  from  the  center  of  axle  to  the 
center  of  rock-shaft,  55  inches ;  length  of  connecting-rod,  84  inches ;  axis  of  cylinder, 
14  inches  above  the  center  of  driving  axle  when  the  engine  is  in  good  working  order. 
We  will  also  assume  that  after  the  driving  axle  has  been  correctly  drawn  in  the 
pedestal,  and  the  rocker  properly  located,  it  is  found  that  the  vertical  distance  from 
the  center  of  axle  to  a  horizontal  line  drawn  through  the  center  of  rock-shaft  is  6 
inches.  It  is  required  to  find,  in  the  order  here  given,  the  off-set  in  the  rocker-arms ; 
the  position  of  crank-pin  for  full  and  half  stroke  of  piston ;  the  radius  of  links ;  the 
position  of  eccentrics ;  the  length  of  eccentric-rods ;  the  position  of  eccentrics  for  half 
strokes  of  piston ;  the  position  of  the  saddle-pin  on  the  links ;  the  position  of  lifting 
shaft  and  the  length  of  the  lifting-shaft  arms  from  which  the  links  are  suspended. 
Indeed,  it  may  be  said  that  to  determine  each  one  of  these  particulars  is  a  problem  by 
itself,  so  that  the  matter  of  laying  out  a  valve  gear  consists  of  the  solutions  of  a  number 
of  simple  problems. 

TO  FIND  THE  OFF-SET  IN  THE    ROCKER-ARMS. 

138.  Let  the  point  A  in  Fig.  200A  be  the  center  of  the  driving  axle ;  through  this 
center  draw  a  vertical  line,  and  make  the  distance  from  A  to  i  equal  to  6  inches; 
through  the  point  i  draw  a  horizontal  line  i  o,  and  make  the  distance  from  i  to  o  equal 
to  55  inches,  as  given  in  the  example ;  the  point  o  will  be  the  center  of  the  rock-shaft. 
From  o,  and  with  a  radius  of  10  inches,  describe  two  arcs,  U  Fand  S  T;  through  the 
center  A  draw  a  line  L  M  tangent  to  the  arc  .S'  7';  this  line  will  be  the  center  line  of 
motion  of  the  valve  gear.  (See  Art.  131.)  Through  the  center  ()  draw  a  line  0  R 
perpendicular  to  L  37;  the  line  O  R  will  be  the  center  line  of  the  lower  arm  of  the 
rocker  when  the  valve  stands  in  the  center  of  its  travel.  Again,  through  0  draw  the 


120 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


vertical  line  0  7;  we  say  vertical,  because  the  valve  is  supposed  to  move  in  a  horizontal 
direction ;  if  the  valve-rod  moves  in  any  other  direction,  then  0  I  should  be  drawn 
perpendicular  to  that  direction.  The  line  0  1  represents  the  center  line  of  the  upper 


J.C 

\                                                                  f 

~-s? 

A\3            JKC 

7- 

c                B                   Bo                 B,    . 

o             (b             n    f 

/                                                                      R 
.X                                           Pig.  200  A. 

T  Re"                                                       "    Fro 
^~~» 

arm  of  rocker  when  the  valve  stands  in  the  center  of  its  travel, 
then  K 11  will  be  the  off-set  in  the  rocker-arms. 


Prolong  I  0  to  K, 


TO  FIND  THE  POSITION   OF  CEANK-PIN  FOE  FULL  AND  HALF   STEOKES   OF  PISTON. 

139.  According  to  the  conditions  given  in  the  example,  the  axis  of  the  cylinder  is 
to  be  1J  inches  above  the  center  A  of  the  axle ;  hence,  on  the  line  A  i  lay  off  a  point  at 
a  distance  of  l£  inches  above  A,  and  through  this  point  draw  the  horizontal  line  E  F, 
which  will  be  the  axis  of  the  cylinder.  From  A  as  a  center,  and  with  a  radius  equal 
to  the  length  of  the  connecting-rod,  namely,  84  inches,  describe  a  short  arc  cutting  E  F 
in  the  point  B2 ;  this  point  will  be  the  center  of  the  crosshead  pin  when  midway  of  its 
travel.  Towards  the  rear  of  B.2  lay  off  the  point  B  at  12  inches  from  B2 ;  also  lay  off 
the  point  B3  towards  the  front  of  B2  at  a  distance  of  12  inches  from  the  latter ;  the 
points  B  and  B3  will  be  the  position  of  the  center  of  the  crosshead  pin  when  the  piston 
is  at  the  extremities  of  a  stroke.  It  should  be  remarked  here  that  the  foregoing  way 
of  finding  the  points  B  and  B3  is  correct  only  when  the  axis  E  F  of  the  cylinder  and 
the  center  A  lie  in  one  and  the  same  straight  line ;  but  when  these  do  not  lie  exactly 
in  the  same  straight  line,  as  is  the  case  in  the  example  before  us,  then  the  construction 
is  not  correct,  but  it  is  sufficiently  close  for  all  practical  purposes,  in  cases  in  which 
the  axis  of  cylinder  is  only  1  £  inches  above  the  center  A  of  axle.  In  exti-eme  cases,  say 
when  the  line  E  F  is  from  3  to  4  inches  above  the  center  A,  it  is  best  to  find  the 
points  B  and  B3  accurately,  which  is  done  in  the  following  manner:  From  A  as  a 
center,  and  with  a  radius  equal  to  the  length  of  the  connecting-rod  minus  the  length  of 
the  crank,  describe  a  short  arc  cutting  E  F  in  the  point  B;  again,  from  A  as  a  center, 
and  with  a  radius  equal  to  the  length  of  the  connecting-rod  plus  the  length  of  the  crank, 
describe  a  short  arc  cutting  E  F  in  the  point  B3 ;  the  distance  from  B  to  B3  will  be 
the  length  of  stroke  of  piston,  which  is  a  little  less  than  twice  the  length  of  crank ;  of 
course  the  length  of  crank  is  the  distance  from  the  center  of  axle  to  the  center  of 
crank-pin.  With  a  radius  of  12  inches  describe  from  A  the  circle  C,  £0,  c,  Jc;  its 
circumference  will  represent  the  path  of  the  center  of  the  crank-pin.  Through  A  and 
B  draw  a  straight  line,  and  prolong  it  to  C  on  the  circumference  of  the  circle ;  this  point 
C  will  be  the  center  of  the  crank-pin  when  the  crosshead  pin  is  at  the  end  B  of  the 
stroke.  Through  A  and  B3  draw  a  straight  line  cutting  the  path  of  the  crank-pin  at  c ; 
this  point  will  be  the  position  of  the  crank-pin  when  the  crosshead  pin  is  at  the  end  B3 


MfiriER\  LOCOMOT/TE    COXSTIIVCTIOX. 


121 


of  tin-  stroke.  For  laying  out  the  valve  gear,  it  is  customary  to  find  the  points  C  and  c 
in  a  somewhat  simpler  way — namely:  through  A  and  7?2  draw  a  straight  line  cutting 
the  path  of  the  '-rank-pin  in  the  points  C  and  c,  which  will  be  the  centers  of  the  crank- 
pin  eorresponding  to  the  extremities  of  the  piston  stroke.  This  way  of  finding  the 
(•••liters  is  not  quite  as  accurate  as  the  first  method  given,  but  the  error  is  so 
small  that  it  may  be  safely  neglected.  In  laying  out  the  valve  gear,  we  have  to  find 
aiiothi-r  two  important  positions  of  the  crank-pin,  namely,  those  corresponding  to  the 
positions  of  the  piston  when  it  is  in  the  center  of  each  stroke,  or,  in  other  words,  when 
tin-  erosshead  pin  is  at  R2.  These  positions  are  found  in  the  following  manner:  From 
B.,  as  a  center,  and  with  a  radius  equal  to  A  R2  (that  is,  the  length  of  the  connecting- 
rod),  descril>e  an  arc  cutting  the  path  of  the  center  of  crank-pin  in  the  points  JC'  and 
£c,  which  will  be  the  required  points.  Join  the  points  A  and  £  (7,  also  A  and  Jc,  by 
straight  lines ;  these  lines  will  be  the  center  lines  of  the  crank  corresponding  to  £  stroke 
of  the  piston ;  and  the  lines  A  C  and  A  c  will  be  the  center  lines  of  the  crank  when  the 
engine  is  on  its  dead  centers.  From  A  as  a,  center,  and  with  a  radius  equal  to  £  of  the 
throw  of  the  eccentric,  2i  inches,  describe  a  circle  cutting  the  center  lines  of  the  crank 
in  the  points  1,  2,  3,  and  4. 

TO  FIND  THE  RADIUS   OP  THE  LINK. 

140.  According  to  construction,  the  point  R  in  Fig.  200A  is  the  center  of  the 
lower  rocker-arm  pin  or  link-block  pin  when  the  valve  stands  midway  of  its  travel ; 
hence,  according  to  Art.  123  the  radius  of  the  link  is  equal  to  the  distance  A  R.    If 
this  radius  is  made  either  longer  or  shorter  than  A  R,  the  tendency  will  be  to  produce 
an  unequal  lead  when  the  link  is  placed  in  mid-gear. 

TO  FIND  THE  POSITIONS  OF  THE  ECCENTRICS  FOR  FULL  STROKES  OF  PISTON. 

141.  This  construction  is  shown  in  Fig.  200B,  which,  for  the  sake  of  clearness,  is 
drawn  to  a  larger  scale  than  Fig.  200 A,  but  in  both  figures  the  lines  and  points  which 
have  the  same  letters  affixed  represent  the  same  parts  of  the  valve  gear. 

On  the  line  L  M  lay  off  from  A  a  distance  A  p  equal  to  the  sum  of  the  lap 


and  lead,  H  inch,  and  through  the  point  /,  draw  a  line  X  7  perpendicular  to  L  M 
cutting  the  circle  1,  ±  :;.  4.  in  th-  points  X  and  Y-.  tin-  j^iut  A",  a«-«-ording  to  Art.  131, 
will  be  the  center  of  the  forward  e<-,-entrie ;  and  Twill  be  the  center  of  the  backward 


122  MODERN  LOCOMOTIVE  CONSTRUCTION. 

eccentric  when  the  crank-pin  is  at  C.  In  this  case  we  have  made  A  p  equal  to  the 
lap  and  lead,  because  the  lengths  of  the  rocker-arms  are  equal.  (See  Art.  131.)  If  the 
lower  rocker-arm  had  been  longer  or  shorter  than  the  upper  one,  then  the  distance 
A  p  would  have  been  made  respectively  greater  or  less  than  the  sum  of  the  lap  and 
lead.  (See  Art.  132.)  On  the  line  L  M  lay  off  from  A  a  distance  A  q  equal  to  A  p, 
and  through  the  point  q  draw  a  line  x  y  parallel  to  X  F,  cutting  the  circumference 
1,  2,  3,  4,  in  the  points  x  and  y.  The  point  x  will  be  the  center  of  the  forward 
eccentric  and  y  that  of  the  backward  eccentric  when  the  crank-pin  is  at  c.  Here,  then, 
we  have  found  the  position  of  eccentrics  when  the  engine  is  on  its  dead  centers. 

TO  FIND  THE  CORRECT  LENGTH  OF  ECC3NTRIC-RODS. 

142.  For  this  purpose  we  shall  need  a  template.      Let  Fig.  200C  represent  the 
link.     Cut  out  a  template  represented  by  the  shaded  portion  of  this  figure.     The  edge 
r  s  of  the  template  must  coincide  with  the  link  arc ;  and  the  arc  t  u  of  the  template 

must  pass  through  the  centers  of  the  eccentric-rod  pins.  Through  the 
centers  of  the  eccentric-rod  pins  t  and  u  draw  straight  lines  on  the  tem- 
plate, and  also  through  a  point  midway  between  these  centers  draw  on  the 
template  a  line  v  w  towards  the  center  from  which  the  link  has  been  drawn  ; 
the  lines  through  t  and  u  may  be  drawn  parallel  to  v  ^v,  or  they  may  also 
point  to  the  center  from  which  the  link  has  been  drawn ;  either  way  will 
answer  the  purpose,  so  long  as  the  points  t,  v,  and  u  on  the  concave  edge 
of  the  template  are  correctly  located  as  explained. 
Pfg.  200  a  From  the  points  X  and  x  as  centers  (Fig.  200B),  and  with  a  radius  equal 

to  A  R  minus  the  width  of  v  w,  draw  two  arcs  X}  and  x}  above  the  line  L  M] 
and  with  the  same  radius  draw  from  the  centers  Fand  y  two  arcs  1",  and  ?/,,  below  L  M. 
Now  place  the  template  with  its  center  line  v  w  on  L  M,  and  with  its  points  t  and  «  on  the 
arcs  X,  and  F,  respectively,  and  along  the  edge  r  s  draw  an  arc  on  the  paper.  Again,  place 
the  template  with  its  center  line  v  w  on  L  M,  and  with  its  points  t  and  u  on  the  arcs  #, 
and  T/J  respectively,  and  along  the  edge  r  s  draw  another  arc  rz  s2.  If  now  these  two 
arcs  which  have  been  drawn  along  r  s  are  at  equal  distance  from  the  point  72,  then 
the  line  drawn  from  X  to  t  or  from  Y  to  u  will  be  the  correct  length  of  the  eccentric- 
rod  ;  but  the  chances  are  that  the  arcs  drawn  along  r  s  will  not  be  at  equal  distances 
from  .R,  because  in  one  position  the  eccentric-rods  will  cross  each  other,  and  in  the 
other  position  they  will  not  do  so,  and  for  these  conditions  we  have  not  made  any 
allowance.  We  must  therefore  draw  another  set  of  arcs  Xj  Yl  and  xl  and  ?/„  with  a 
somewhat  larger  radius,  and  again  place  the  concave  side  of  the  template  on  these  arcs 
and  draw  arcs  along  the  edge  r  s  as  before.  With  a  little  care  and  good  judgment  in 
drawing  Xt,  F];  x^  and  ylt  the  arcs  drawn  along  the  edge  r  s  of  the  template  will  now 
be  at  equal  distances  from  the  point  jR,  and  therefore  the  lines  joining  the  points  X 
and  t  or  Y  and  u  will  be  the  correct  length  of  eccentric-rods. 

143.  The  heavy  lines  X  t  and  Y  u  represent  the  position  of  the  center  lines  of  the 
eccentric-rods  when  the  crank-pin  is  at  (7;  it  is  seen  that  these  rods  cross  each  other; 
but  when  the  crank-pin  is  at  c  the  eccentric-rods  will  not  cross  each  other ;  they  are 
then  said  to  be  "  open."    If  we  connect  the  rods  so  that  they  will  not  cross  each  other 


MODERN   LOCOMOTIVE    CONSTRUCTION. 


123 


when  tin-  crank-pin  is  at  (7,  we  shall  have  no  lead  when  the  link  is  placed  in  mid-gear, 
and  this  is  not  admissible  iu  locomotive  practice,  although  it  may  be  advantageously 
employed  in  hoisting  engines,  because  when  there  is  no  lead  with  the  link  in  mid-gear 
the  engine  can  be  stopped  simply  by  raising  the  link  without  touching  the  throttle- 
valve. 

TO   FIND  THE  POSITION   OF  ECCENTRICS   FOR  HALF   STROKES   OF  PISTON. 

144.  To  show  this  construction  we  shall  refer  to  Fig.  200D,  to  which  nearly  all  the 
lines  shown  in  the  former  figure  have  been  transferred.  We  have  seen  that  the  center 
lines  of  the  crank  corresponding  to  full  and  half  stroke  of  piston  cut  the  path  of  the 
center  of  eccentrics  in  the  points  1,  2,  3,  and  4.  Now  when  the  crank-pin  is  at  C 


the  forward  eccentric  will  be  at  X,  and  the  length  of  the  arc  from  the  point  1  to  X  is 
fixed ;  it  cannot  be  changed  in  whatever  position  the  crank  may  be ;  hence,  when  the 
crank-pin  is  at  J  6'  the  arc  from  the  point  2  to  £X  must  be  equal  to  that  from  1  to  X. 
Therefore,  making  the  arc  JX-2  equal  to  X-l,  we  obtain  the  point  £X,  which  is  the 
center  of  the  forward  eccentric  when  the  piston  is  in  the  center  of  the  forward  stroke. 
For  similar  reasons  we  make  the  arc  Ja>4  equal  to  X-l ;  the  point  ^x  will  be  the  center 
of  the  forward  eccentric  when  the  piston  is  in  the  center  of  its  return  stroke.  Of  course, 
care  must  be  taken  to  lay  these  points  off  on  the  path  of  the  center  of  eccentrics,  that  is, 
on  the  circumference  X  Y  x  y,  and  these  points  should  be  laid  off  to  follow  the  crank. 
We  know  that  the  point  Y  is  the  center  of  the  backward  eccentric  when  the  crank  is 
at  C,  and  the  length  of  the  arc  from  1  to  Y  remains  fixed  for  all  positions  of  the  crank ; 
therefore  make  the  arc  %Y-2  equal  to  Y-i,  also  make  the  arc  £«/— 4  equal  to  the  same 
arc;  of  course,  the  points  ^Y  and  %y  must  be  ahead  of  the  crank.  Now  the  point  %Y 
will  be  the  position  of  the  backward  eccentric  when  the  piston  is  in  the  middle  of  its 
forward  stroke,  and  the  point  \y  will  be  the  position  of  the  backward  eccentric  when 
the  piston  is  in  the  middle  of  its  return  stroke. 


TO   FIND   THE   CENTER   OF  THE   SADDLE-PIN. 


145.  From  the  centers  iX  and  4.r,  ami  with  a  radius  equal  to  X  t  or  Y u  (Fig.  200B), 
describe  in  Fig.  200D  above  the  line  L  3/the  arcs  i  A,  and  i./, ;  from  the  centers  A  I'aml 


124  MODERN  LOCOMOTIVE   CONSTRUCTION. 

%y,  and  with  the  same  radius,  describe  below  the  line  L  M  the  arcs  £  Y}  and  %y}.  It  will 
be  remembered  that  the  point  E  is  the  center  of  the  link-block  pin  when  the  valve 
stands  midway  of  its  travel ;  from  this  point  E  as  a  center,  and  with  a  radius  equal  to 
the  lap  of  valve  £  of  an  inch,  describe  the  lap  circle,  cutting  the  line  L  M  in  the  points 
a  and  b.  Now  place  the  template  with  its  point  t  on  the  arc  £X,,  and  with  the  point 
u  on  the  arc  &YU  and  its  outer  edge  r  s  touching  the  point  a;  in  this  position  draw  on 
the  paper  along  the  outer  edge  of  the  template  an  arc  r  s,  and  also  mark  off  the  points 
v  and  w ;  lift  off  the  template,  and  through  the  points  v  and  w  draw  a  straight  line  v  w ; 
this  line  represents  the  position  of  the  center  line  of  the  link  when  steam  is  being  cut 
off  at  J  of  the  forward  stroke.  Again,  lay  the  template  in  a  position  as  indicated  by 
the  darker  shading;  the  point  t  must  lie  on  the  arc  \x-±;  the  point  «,  on  the  arc  £?/, ; 
and  its  outer  edge  r2  s2  must  touch  the  point  b.  In  this  position  draw  on  the  paper 
an  arc  r2  s2  along  the  outer  edge  of  the  template,  and  also  mark  off  the  points  v2  wz ; 
then  lift  off  the  template  and  join  the  points  v2  w2  by  a  straight  line ;  this  line  will  be 
the  center  line  of  link  when  steam  is  cut  off  at  \  of  the  return  stroke.  On  the  line 
v  w  lay  off  a  point  f,  and  on  the  line  v.,  ivz  lay  off  a  point  /2,  to  comply  with  the  follow- 
ing conditions :  The  distance  from  the  point  f  to  the  arc  r  s  must  be  equal  to  the 
distance  from  the  point  f,  to  the  arc  r2  s2,  and  these  points  f,  f,  must  also  lie  in  a  line 
parallel  to  L  M ;  the  position  of  these  points  are  found  by  trial — usually  two  or  three 
trials  will  locate  them  correctly.  The  distance  from  the  arc  r  s  to  the  point  f, 
or  the  distance  from  the  arc  r2  s2  to  the  point  f2  (these  distances  being  equal),  will  be 
the  amount  that  the  saddle-pin  has  to  be  set  inwards  from  the  link  arc.  Mark  off 
correctly  this  point  /on  the  template. 

TO   LOCATE   THE   LIFTING   SHAFT. 

146.  On  the  line  L  M  lay  off  from  E  a  point  c2 ;  the  distance  from  E  to  c2  must  be 
equal  to  the  sum  of  the  lap  and  lead  If  inches.  Also  from  E  lay  off  on  L  M  a,  point  c3 ; 
the  distance  from  E  to  c.j  must  again  be  equal  to  the  sum  of  the  lap  and  lead.  Let  it 
now  be  remembered  that  the  arcs  Xj,  F1?  xl  and  yl  in  Fig.  200D  represent  the  same  arcs 
as  those  which  have  the  same  letters  affixed  in  Fig.  200B.  Now  place  the  template  as 
low  as  possible,  with  its  point  t  on  the  arc  X19  the  point  u  on  the  arc  Yly  and  the  outer 
edge  r  s  touching  the  point  c2.  In  this  position  prick  off  through  the  point  /  on  the 
template  the  point  d  on  the  paper.  Now  move  the  template  to  as  low  a  position  as 
possible,  so  that  its  outer  edge  r  s  will  touch  the  point  c:j  with  the  point  t  on  the  arc  #„ 
and  the  point  u  on  the  arc  ylt  and  in  this  position  prick  off  through  /on  the  template 
a  point  e  on  the  paper.  From  the  points  d  and  e  as  centers,  and  with  a  radius  equal  to 
the  length  of  the  link-hanger,  13  inches,  describe  two  arcs  intersecting  each  other  in  the 
point  k.  Now  move  the  template  as  high  as  possible,  and  let  its  outer  edge  r  s  touch 
the  point  c2,  and  let  the  point  t  be  on  the  arc  X1?  and  the  point  u  on  the  arc  Ylt  then 
through  the  point /on  the  template  prick  off  on  the  paper  a  point  g.  Again,  move  the 
template  to  as  high  a  position  as  possible,  and  let  its  outer  edge  r  s  touch  the  point  r3, 
and  let  the  point  t  be  on  the  arc  x^  and  u  on  the  arc  ?/, ;  through  /  on  the  template 
prick  off  on  the  paper  a  point  7).  From  the  points  g  and  h  as  centers,  and  with  a 
radius  equal  to  the  length  of  the  link-hanger,  describe  two  arcs  intersecting  each  other 


3TODKRX  LOCOMOTIVE   COXSTRVCTIOX. 

in  the  point  n.  Lastly,  from  tho  points  /  and  f.,  previously  located  on  the  paper, 
describe  with  a  radius  of  13  inches  two  arcs  intersecting  each  other  in  the  point  I. 
Now  find  by  trial  a  point  m  from  which  an  arc  passing  through  the  points  kin  can 
lie  described;  this  point  »/  will  be  the  center  of  the  lifting  shaft,  and  the  radius  m  n 
will  be  the  length  of  the  lifting-shaft  arms  from  which  the  links  are  to  be  suspended. 

This  completes  the  whole  construction.  It  will  be  noticed  that  in  this  valve 
motion  the  valve  will  have  equal  lead  at  full  stroke,  and  cut  off  steam  at  the  center  of 
the  forward  and  return  stroke.  Under  these  conditions  steam  will  be  cut  off  at  equal 
portions  of  the  forward  and  return  strokes,  or  very  nearly  so,  for  all  intermediate 
positions  of  the  liuk. 

147.  lu  designing  an  engine  the  foregoing  construction  is  essential ;  it  gives  us 
the  required  data  for  several  details  which  can  be  made  and  completely  finished  before 
they  are  sent  into  the  erecting  shop.  But  there  are  other  details  which  have  to  be 
adjusted  to  the  imperfect  workmanship  which  cannot  always  be  avoided;  for  instance, 
the  length  of  eccentric-rods  will  have  to  be  adjusted  under  the  engine,  also  the 
eccentrics  will  have  to  be  set  on  the  axle,  and  there  are  other  points  which  cannot  be 
exactly  transferred  from  the  paper  to  the  engine ;  it  will  therefore  be  advantageous  to 
examine  now  the  mode  of  procedure  in  setting  the  valves  on  the  engine  in  the  shop. 

PEACTICAL  EXAMPLE  OF  SETTING  THE  VALVE  GEAK  OF  A  LOCOMOTIVE  IN   THE   ERECTING  SHOP. 

EXAMPLE  43. — Throw  of  the  eccentric  is  5  inches;  lap,  •}-§•  of  an  inch;  lead, 
-j1,,  of  an  inch ;  clearance  between  piston  and  each  cylinder  head  is  j}  of  an  inch ;  type 
of  locomotive  is  one  in  which  the  eccentrics  are  placed  on  the  main  axle,  that  is,  the 
axle  to  which  the  main  rod  is  connected :  it  is  required  to  set  or  adjust  the  valve  gear. 
In  this  example  the  position  of  the  saddle-pin  on  the  saddle  is  assumed  to  be  correct, 
so  that  in  setting  the  valve  gear  no  loose  saddle  is  to  be  used. 

To  the  young  mechanic,  it  may  appear  that  setting  the  valve  gear  of  a  locomotive 
is  a  very  mysterious  and  difficult  operation.  Yet,  if  he  understands  the  theory  of  the 
valve  motion,  which  we  have  endeavored  to  explain  and  to  make  plain  in  the  foregoing 
.•irticles,  and  proceeds  in  setting  the  valve  gear  in  a  systematic  manner,  this  opera- 
tion will  lose  all  its  mystery,  and  will  not  be  such  a  difficult  problem  as  it  first 
appeared  to  be.  True,  in  attempting  to  set  the  valve  gear  we  are  confronted  with 
that  which  may  seem  to  be  a  very  mixed-up  problem,  because  the  correct  position  of 
the  eccentrics  on  the  axle  and  the  exact  length  of  the  eccentric-rods  must  be 
determined ;  and  since  these  two  items  are  intimately  connected,  we  may  often  be 
in  doubt  as  to  which  of  the  two  has  been  correctly  found.  But  by  dividing  this 
problem  into  a  number  of  simpler  ones,  its  solution  will  be  comparatively  easy.  In 
connection  with  the  setting  of  the  valve  gear  a  few  preliminary  preparations  are 
necessary;  this  whole  subject  will  be  treated  in  the  following  order: 

1st.  Blocking  up  the  main  axle  boxes  in  the  pedestals  of  the  frames. 

2d.  To  find  the  exact  position  of  tho  crank-pin  when  the  piston  is  at  the  beginning 
of  a  stroke. 

3d.  To  lest  the  amount  of  clearance  between  the  piston  and  cylinder  heads. 

4th.  To  connect  the  eccentric-rods  correctly  to  the  liuk. 


126  MODERN  LOCOMOTIVE   CONSTRUCTION. 

5th.  To  find  the  correct  length  of  the  eccentric-rods. 

6th.  To  find  the  correct  position  of  the  eccentrics  on  the  axle. 

7th.  To  lay  off  the  notches  in  the  reverse  arcs  or  quadrants. 

In  this  example  it  is  assumed  that  the  quadrants  are  properly  fastened  in  their 
position,  and  that  they  have  no  notches.  Also  that  the  reverse  lever  and  lifting 
shaft  are  in  correct  positions ;  and  that  the  reach-rod  connects  the  reverse  lever  and 
lifting  shaft ;  and  lastly,  that  the  links  are  suspended  from  the  lifting-shaft  arms,  and 
that  the  former  are  properly  attached  to  the  lower  rocker-arms;  and  also,  that  the 
connections  between  the  upper  rocker-arms  and  the  valves  are  complete. 

BLOCKING  UP  THE  MAIN  AXLE  BOXES  IN  THE  PEDESTALS  OF  THE  FKAMES. 

148.  The  locomotive  should  rest   securely  on  blocks,  so  that  the  main  driving 
wheels  can  be  lifted  off  the  track  and  held  in  a  position  which  will  allow  the  main 
driving  wheels  to  be  easily  turned  around  in  their  boxes.     In  lifting  the  main  driving 
wheels  off  the  track,  they  should  be  raised  to  a  position  in  the  pedestal  which  corre- 
sponds to  the  position  of  their  boxes  in  the  frame  when  the  engine  is  hauling  a  train. 

The  general  practice  is  to  give  l£  inches  clearance  between  the  axle  box  and  pedestal 
cap,  and  about  3  inches  clearance  between  the  top  of  axle  box  and  frame.  In  smaller 
engines  whose  axle  boxes  have  less  clearance  in  the  pedestals,  the  proportion  between 
the  upper  and  lower  clearances  should  be  about  the  same  as  that  just  given ;  namely, 
the  clearance  on  top  of  the  axle  box  should  be  about  twice  as  great  as  that  below  the 
box.  (See  Figs.  271  to  279.)  If  then  1J  inches  is  to  be  the  clearance  at  the  bottom, 
blocks  each  l£  inches  thick  are  inserted  between  the  main  axle  boxes  and  the  pedestal 
caps,  which  will  hold  the  main  driving  wheels  off  the  track  and  in  the  correct  position 
during  the  time  the  valve  motion  is  to  be  adjusted.  No  attention  need  be  paid  to 
the  other  boxes  and  wheels,  since  in  this  class  of  engines  it  is  not  necessary  to  put  on 
the  side  rods  for  the  purpose  of  setting  the  valve  gear ;  in  fact,  the  side  rods  would  be 
in  the  way,  and  would  give  unnecessary  labor  in  revolving  the  main  driving  wheels. 

TO  FIND  THE  EXACT  POSITION  OF  THE  CRANK  WHEN  THE  PISTON  IS  AT  THE  BEGINNING  OF 
A  STROKE;    OR,  IN  OTHER  WORDS,  TO  LOCATE  THE  DEAD  CENTERS  OF  THE  CRANK. 

149.  Tig.  201.     The  following  remarks  refer  only  to  the  setting  of  the  valve  gear 
on  the  right-hand  side  of  the  engine ;  the  manner  of  setting  the  valve  gear  on  the 
left-hand  side  of  the  engine  is  precisely  the  same  as  that  adopted  for  the  right-hand 
side.     First  set  or  adjust  one  side,  and  then  the  other  side.     On  the  face  of  the  tire 
describe  with  the  aid  of  a  gauge  two  arcs  d  e  and  /  g  in  about  the  same  relative 
positions  to  that  of  the  crank,  as  shown  in  the  figure ;  care  should  be  taken  to  describe 
these  arcs  as  true  as  if  they  had  been  described  in  the  lathe.     Attach  the  main  rod  to 
the  crank-pin  and  crosshead ;  but  at  present  do  not  fasten  the  piston-rod  to  the  cross- 
head.    Now  turn'  the  driving  wheel  in  the  direction  of  the  arrow  until  the  crosshead 
is  within  a  short  distance  from  the  end  of  the  stroke,  say  \  an  inch;  while  in  this 
position  mark  on  the  slides  a  line  i  even  with  the  end,/  of  the  crosshead.    Also,  while 
the  wheel  and  crosshead  are  in  this  position,  place  a  center  punch  mark  c  on  the  side 


LOCOMOTICK    COXSTKI:CI'I<>.\. 


127 


of  the  fire-box  or  any  other  convenient 
fixed  piece  of  mechanism.     From  this 
point  c  as  a  center,  and  with  a  tram  of 
any  convenient  length  as  a  radius,  de- 
srribe  on  the  face  of  the  tire  a  short 
a  iv,   cutting  the  arc  d  e  (previously 
marked  on  the  tire)  at  the  point  d  (this 
point  d  will  at  this  instant  coincide 
with  the  end  h  of  the  tram,  and  not  as 
shown  in  the  figure).    Now  turn  the 
wheel  in  the  same  direction  as  before, 
causing  the  crosshead  to  complete  its 
full  stroke  and  a  very  small  portion  of 
its  return  stroke;   and  when   during 
this  motion  the   edge  j  of  the  cross- 
head  touches  the  lino  i  on  the  slides, 
stop  turning  the  shaft,  and  while  in 
this  position  describe  from  the  center 
punch  mark  c  as  a  center,  and  with  the 
same  tram  as  before,  a  short  arc  on 
the  face  of  the  tire,  cutting  the  arc  d  c 
at  the  point  e.    Midway  between  the 
points  d  and  e  on  the  arc  d  e  lay  off 
point  It.    Turn  the  wheel  in 
a  position  in  which  the  tram 
will  touch  the  points  c  and 
h  •  the  crank-pin  will  then 
be  at  B,  and  the  piston  will 
be.  at   the  beginning  of  its 
forward  stroke ;  or,  in  other 
words,  the  crank  will  then 
be    on    one    of    its     dead 
centers.     While  the  crank 
stands   at   J>    draw  on   the 
slides  a  line  /,•  even  with  the 
end  j  of  the  crosshead ;  this 
line  k  will  indicate  the  end 
of  the  stroke — that  is,  the 
end  ./  of  the  crosshead 
will  not  travel  beyond 
the  line  /,•  with  the  con- 
necting-rod    attached. 
In  order  to  locate  the 
other  dend  center  .1  wo 
again  turn  the  wheel  in 


128  MODERN  LOCOMOTIVE   CONSTRUCTION. 

the  direction  of  the  arrow  until  the  crosshead  is  within  a  short  distance  of  the  front 
end  of  the  stroke,  say  £  an  inch,  and  while  in  this  position  mark  on  the  slides  a 
line  m  even  with  the  end  n  of  the  crosshead ;  also,  before  the  wheel  and  crosshead  are 
moved  out  of  this  position,  describe  on  the  face  of  the  tire  from  the  point  c  as  a  center, 
and  with  the  same  tram  as  used  before,  a  short  arc,  cutting  the  arc  fg  at  the  point  g. 
Next  turn  the  wheel  in  the  same  direction  as  before,  causing  the  crosshead  to  com- 
plete its  forward  stroke  and  a  small  portion  of  its  backward  stroke,  and  when  the  edge 
n  of  the  crosshead  again  touches  the  line  m  on  the  slides,  stop  turning  the  wheel. 
From  the  center  c,  and  with  the  same  tram,  describe  on  the  face  of  the  tire  a  short 
arc,  cutting  the  arc  fg  at  the  point/  Mark  off  the  point  I  on  the  arc  fg  midway 
between  the  points  /and  g.  Now,  turning  the  wheel  in  a  position  in  which  the  tram 
will  touch  the  points  c  and  I,  the  center  of  the  crank-pin  will  be  at  A,  which  is  the 
other  dead  center.  While  the  crank-pin  is  at  A  draw  on  the  slides  a  line  o  even  with 
the  end  n  of  the  crosshead ;  this  line  o  will  indicate  the  other  end  of  the  stroke. 

TO  TEST   THE  AMOUNT  OF  CLEARANCE  BETWEEN  THE  PISTON  AND  CYLINDER  HEADS. 

150.  For  this  purpose  we  must  take  off  the  main  rod,  fasten  the  piston  to  the 
piston-rod,  key  the  piston-rod  to  the  crosshead,  and  bolt  the  cylinder  heads  to  the 
cylinder.  Our  example  calls  for  f  of  an  inch  clearance  between  piston  and  cylinder 
head  at  each  end  of  the  cylinder.  Consequently  the  crosshead  with  piston-rod  and 
piston  must  be  pulled  out  as  far  as  can  be  done ;  if  then  the  edge  j  of  the  crosshead 
overlaps  the  line  k,  $  of  an  inch,  the  clearance  is  correct ;  if  the  edge  j  of  the  crosshead 
only  reaches  £  of  an  inch  beyond  the  line  &,  then  the  clearance  is  £  of  an  inch  too 
small ;  or  if  the  edge  j  of  the  crosshead  does  not  reach  the  line  k,  then  there  is  no 
clearance,  and  the  cylinder  heads  are  liable  to  be  broken  if  we  attempt  to  turn  the 
crank  over  the  dead  centers.  In  a  similar  manner  we  find  the  amount  of  clearance  at 
the  front  end  of  the  cylinder ;  that  is,  we  push  the  crosshead  towards  the  front  as  far 
as  it  will  go,  then,  if  the  edge  n  of  the  crosshead  moves  §  of  an  inch  beyond  the  line  o, 
the  clearance  is  correct;  if  less,  the  clearance  is  insufficient.  If  the  clearance  is 
insufficient  at  the  back  end  of  the  cylinder,  then  a  corresponding  amount  must  be 
turned  off  the  back  cylinder  head.  An  insufficient  clearance  at  the  front  end  may  be 
caused  by  the  heads  of  the  follower  bolts  projecting  too  far  beyond  the  piston ;  in  that 
case  a  groove  must  be  turned  in  the  front  cylinder  heads,  to  clear  the  heads  of  the 
follower  bolts ;  or  the  insufficient  clearance  may  be  caused  by  the  front  cylinder  head 
projecting  too  far  into  the  cylinder ;  in  this  case  a  sufficient  amount  must  be  turned 
off  the  front  cylinder  head.  In  some  cases  an  insufficient  clearance  may  be  caused  by 
the  piston  being  fitted  out  of  square  on  the  piston-rod ;  in  that  case  the  faces  of  the 
piston  must  be  turned  off.  The  writer  in  his  experience  has  seldom  found  the  clearance 
to  be  too  great,  but  has  often  found  it  to  be  too  small,  generally  owing  to  the  rough- 
ness of  the  faces  of  the  piston  and  inside  surface  of  the  cylinder  heads,  as  these  are 
seldom  turned  unless  it  is  absolutely  necessary  to  do  so  for  the  correct  amount  of 
clearance. 


M<H>i:i;.\   l.ornMilTlfK   CONSTRUCTION. 


TO  CONNECT  THE  ECCENTRIC-RODS  CORRECTLY  TO  THE  LINK. 

151.  In  locomotives  it  is  necessary  to  preserve  the  lead,  no  matter  in  what  position 
the  link  may  be  placed.     In  Art.  108  we  have  shown  that  the  eccentric-rods  can  be 
connected  to  the  link  in  a  manner  which  will  increase  the  lead  as  the  link  is  moved 
from  full-gear  to  mid-gear.    In  Art.  110  it  is  also  shown  that  the  eccentric-rods  can  be 
connected  to  the  link  in  a  manner  which  will  give  lap  instead  of  lead  when  the  link  is 
in  mid-gear.    Now,  since  we  must  have  lead  in  whatever  position  the  link  may  be 
placed,  it  follows  that  the  eccentric-rods  must  be  connected  to  the  link  in  a  manner 
which  will  accomplish  the  desired  result.     In  Fig.  201  the  piston  is  at  the  beginning  of 
its  forward  stroke,  and  the  eccentric-rods  are  shown  to  cross  each  other,  and,  conse- 
quently, when  the  axle  has  made  one-half  of  a  revolution,  and  the  piston  is  at  the  begin- 
ning of  the  backward  stroke,  the  eccentric-rods  will  not  cross  each  other,  as  shown  in 
Fig.  29  (Art.  52).     This  manner  of  connecting  the  eccentric-rods  to  the  link  is  correct, 
because  it  will  give  lead,  no  matter  in  what  position  the  link  is  placed,  although  in  mid- 
gear  the  lead  will  be  greater  than  in  full-gear.     (See  Art.  108.)     The  reason  why  the 
manner  of  connecting  the  eccentric-rods  to  the  link  should  affect  ^the  lead  is  clearly 
shown  in  Fig.  202.     In  this  figure  the  different  parts  of  the  valve  gear  are  represented 
by  center  lines  ;  the  arc  d  I  represents  an  arc  drawn  through  the  center  of  the  link 
opening,  and  the  arc  e  g  represents  the  arc  in  which  the  centers  of  eccentric-rod  pins 
arc  located.    The  full  lines  show  the  correct  manner  of  connecting  the  eccentric-rods 
to  the  link,  and  the  dotted  lines  represent  the  incorrect  manner  of  doing  this.    Here 
it  will  be  seen  that  by  connecting  the  eccentric-rods,  as  shown  by  the  full  lines,  we 
have  lead  when  the  link  is  placed  in  mid-gear  ;  and  by  changing  the  connections,  as 
shown  in  dotted  lines,  we  push  the  lower  rocker-arm  pin  away  from  the  axle,  which 
causes  the  upper  rocker-arm  pin  to  pull  the  valve  towards  the  axle,  and  thus  close  the 
port  and  destroy  the  lead  when  the  link  is  in  mid-gear.     In  connecting  the  eccentric- 
rods  to  the  link,  as  shown  in  Fig.  201,  we  must  not  forget  that  a  rocker  is  employed. 
If  a  rocker  had  not  been  interposed,  then  the  eccentric-rods  would  not  cross  each 
other  when  the  piston  is  at  the  beginning  of  the  forward  stroke,  but  would  be  open,  as 
shown  in  Fig.  1G8  (Art.  108). 

TO  FIND  THE  CORRECT  LENGTH   OF  THE  ECCENTRIC-RODS. 

152.  Fig.  203,  to  which  we  now  shall  refer,  is  supposed  to  represent  the  same  valve 
gear  as  that  shown  in  Fig.  201.     But  for  the  sake  of  clearness  we  have  left  out  all 
details  not  necessary  for  a  clear  conception  of  the  subject.     For  instance,  instead  of 
showing  the  driving  wheel  with  arms  and  crank,  as  has  been  done  in  Fig.  201,  we  have 
simply  shown  the  tire  of  the  same  driving  wheel  with  the  arcs  <i  h  c  and////  marked 
ii]  ton  its  face.     Of  course,  these  arcs  represent  the  same  arcs  as  shown  on  the  face  of 
tire  in  Fig.  201.     In  Fig.  203  the  center  line  of  crank  is  represented  by  the  straight  line 
B  6'  or  A  C;  B  being  the  center  of  crank-pin  when  it  is  on  the  rear  dead  center,  and 
A  the  center  of  crank-pin  when  on  the  front  dead  center;   (!  is  the  center  of  the  driving 
axle.     The  circumference  x  //  represents  the  path  of  the  centers  of  eccentrics.     The 
distance  between  the  rocker  and  the  slide-valve  has  been  shortened  and  the  slide-valve 


130 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


I 


drawn  to  a  larger  scale  than  the  other  details.  The 
slides  have  been  left  out,  because  for  the  present 
purpose  no  attention  need  be  paid  to  them.  The 
tram  c  h,  the  reverse  lever  H,  the  quadrant  G, 
reach  rod,  and  lifting  shaft  represent  the  same 
details  as  those  shown  in  Fig.  201.  Also,  the  cen- 
ter punch  marks  c  in  Figs.  201  and  203  represent 
one  and  the  same  point.  Drawing  a  part  of  Fig. 
203  to  a  larger  scale  than  the  other  part  will  not 
affect  the  correctness  of  the  following  reasoning 
and  explanation : 

In  order  to  save  as  much  time  and  labor  as 
possible  when  the  correct  length  of  each  eccentric- 
rod  is  to  be  determined,  it  will  be  best  to  obtain 
from  the  drawing-room  the  correct  relative  posi- 
tion of  the  eccentrics  to  that  of  the  crank,  also  the 
length  of  the  eccentric-rods ;  and  then  first  set  the 
eccentrics  on  the  driving  axle  in  a  position  which 
appears  to  be  sufficiently  accurate,  or,  in  other 
words,  set  the  eccentrics  on  the  axle  as  nearly  cor- 
rect as  can  be  done  without  spending  too  much 
time,  or  making  special  measure- 
ments. Also  adjust  the  length 
of  the  eccentric-rods  to  the  length 
obtained  from  the  drawing-room, 
and  bolt  the  rods  to  the  eccentric- 
strap,  and  connect  the  rods  to 
the  link  as  previously  explained. 
Now,  the  eccentric-rods  being 
properly  connected  to  the  link, 
our  next  step  will  be  to  find  the 
correct  elevation  or  location  of 
the  link,  and  the  position  of  the 
reverse  lever  //  when  the  valve 
motion  is  in  full-gear  forward, 
and  also  in  full-gear  backward. 
The  general  practice  is  to  suffi- 
ciently raise  or  lower  the  link 
which  will  not  allow  the  link- 
block  to  approach  the  ends  of  the 
slot  in  the  link  closer  than  £  or  jj 
of  an  inch  during  any  one  stroke. 
Hence,  we  may  say  the  least  clear- 
ance between  the  link-block  and 
the  link  is  from  f  to  £  an  inch. 


MODERN  LOCO.VOTIJ'E  CONSTRUCTION. 

us  assume  that  this  clearance  is  to  be  %  an  inch.  Hence,  turn  the  driving  wheel  in 
the  direction  of  the  arrow  1  (Fig.  203),  and  in  the  meantime  move  the  reverse  lever  H into 
the  forward  gear,  and  when  the  lever  is  in  a  position  which  will  cause  the  least  clear- 
ance between  the  top  of  the  link-block  and  the  end  of  the  slot  in  the  link,  during  one 
revolution  of  the  wheel,  to  be  equal  to  £  an  inch,  stop  turning  the  wheel  and  clamp 
the  lever  //  to  the  quadrants  G.  This  location  of  the  link  will  be  that  to  which  it 
must  be  lowered  when  the  valve  motion  is  in  full-gear  forward ;  and  since  the  elevation 
or  location  of  the  link  is  regulated  by  the  reverse  lever  //,  it  follows  that  the  position  of 
the  lever  H  as  found  is  for  full-gear  forward.  Therefore  on  the  quadrants  G  mark  this 
position  of  the  reverse  lever  //,  so  that  at  any  time  afterwards  it  can  again  be  readily 
placed  in  the  same  position.  Next  turn  the  wheel  in  a  direction  opposite  to  that  of 
the  arrow,  and  in  the  meantime  move  the  reverse  lever  H  towards  the  rear  end  of  the 
quadrants  G,  and  when  the  link  is  sufficiently  raised  so  as  to  cause  the  least  clearance 
between  the  bottom  end  of  the  link-block  and  end  of  the  slot  in  the  link  to  be  equal  to 
%  an  inch  during  one  revolution  of  the  wheel,  stop  turning  the  same,  clamp  the  lever 
//to  the  quadrants  G,  and  mark  its  position  on  the  quadrants  for  future  use.  Now,  to 
determine  the  length  of  the  eccentric-rods,  we  simply  lengthen  or  shorten  the  rods  so 
that  the  distance  between  the  forward  end  s  of  the  extreme  travel  of  the  slide-valve 
and  the  outer  edge  r  of  the  front  steam  port  will  be  equal  to  that  between  the  rear  end 
s2  of  the  extreme  travel  of  the  valve  and  the  outer  edge  t  of  the  rear  steam  port.  In 
adjusting  the  length  of  the  eccentric-rods,  we  may  commence  with  the  forward  or  back- 
ward eccentric-rods ;  let  us  commence  with  the  former.  Place  the  reverse  lever  H — 
and  thus  the  whole  valve  motion — into  the  full  forward  gear,  and  clamp  the  lever  H  to 
the  quadrants  G.  Place  a  small  ordinary  square  against  the  edge  p  of  the  slide-valve ; 
then  turn  the  driving  wheel  in  the  direction  of  the  arrow  1,  causing  the  slide-valve  to 
move  in  the  direction  of  the  arrow  marked  2,  and  pushing  the  small  square  before  it. 
When  the  valve  has  reached  its  extreme  end  s  of  travel,  the  valve  will  commence  to 
move  in  an  opposite  direction,  as  indicated  by  the  arrow  3,  and  the  square  will  be  left 
standing,  indicating  the  distance  between  the  outer  edge  r  of  the  front  steam  port  and 
the  extreme  end  s  of  the  travel  of  the  valve.  "While  the  square  is  in  this  position  draw 
on  the  steam-chest  seat  a  line  touching  the  edge  s  of  the  square.  Next  place  the 
squar<»  against  the  edge  o  of  the  valve,  and  continue  turning  the  driving  wheel  in  the 
same  direction  as  before,  causing  the  valve  to  travel  in  the  direction  of  the  arrow  3, 
and  pushing  the  square  towards  the  rear  end  s.,  of  the  travel.  When  the  valve  has 
pushed  the  square  towards  the  rear  as  far  as  it  can  do,  draw  on  the  steam-chest  seat  a 
line  touching  the  edge  s2  of  the  square;  this  line  s2  indicates  the  rear  end  of  the 
extreme  travel  of  the  valve,  and  the  distance  s2  t  shows  the  amount  of  travel  of  the 
valve  beyond  the  outer  edge  t  of  the  rear  steam  port.  If  the  distance  s2  t  is  equal  to 
the  distance  r  s,  the  length  of  forward  eccentric-rod  is  for  the  present  considered  to  be 
correct.  If  r  sis  greater  than  s2  t,  the  eeeontrie-rod  is  too  short  and  must  be  lengthened ; 
if  r  s  is  less  than  s.2  t,  the  eccentric-rod  must  be  made  shorter.  Now  place  the  reverse 
lever  /JTinto  full-gear  backward,  and  damp  it  to  the  quadrants  G.  Turn  the  wheel 
in  a  direction  opposite  to  that  of  the  arrow  1,  and  with  the  aid  of  the  small  square  find 
tit--  extreme  travel  of  the  slide-valve  in  a  manner  precisely  the  same  as  that  just 
explained.  And  if  in  this  case  the  distances  s.,  t  and  r  s  are  equal,  the  length  of  the 


132  MODERN  LOCOMOTIVE   CONSTRUCTION. 

backward  eccentric-rod  is  for  the  present  considered  to  be  correct.  If  these  distances 
s2  t  and  r  s  are  not  equal,  the  backward  eccentric-rod  must  be  lengthened  or  shortened 
as  the  case  may  require.  If  the  backward  eccentric-rod  has  to  be  lengthened  or 
shortened,  the  valve  motion  should  again  be  set  into  full-gear  forward  and  examined, 
so  as  to  determine  that  the  adjustment  of  the  length  of  the  backward  eccentric  has  not 
affected  the  action  of  the  forward  eccentric-rod.  It  is  possible  that  the  length  of  the 
forward  eccentric-rod  may  want  a  little  readjustment. 


TO   FIND  THE  CORRECT  POSITION   OF  THE  ECCENTRICS   ON  THE  AXLE. 

153.  Place  the  reverse  lever  H  into  full-gear  forward ;  turn  the  driving  wheel  in 
the  direction  of  the  arrow  1,  Fig.  203,  till  the  tram  will  touch  the  center  punch  marks 
c  and  h,  and  then  stop  turning  the  wheel.  When  the  wheel  is  in  this  position,  the 
crank-pin  will  be  on  the  rear  dead  center  B.  Move  the  forward  eccentric  around  the 
driving  axle  until  the  required  lead  is  obtained — that  is,  until  the  valve  has  opened  the 
rear  steam  port  ia6  of  an  inch.  Again,  turn  the  driving  wheel  in  the  same  direction  as 
before,  till  the  tram  will  touch  the  center  punch  marks  c  and  I ;  then  stop  turning  the 
wheel.  While  the  wheel  is  in  this  position,  the  crank-pin  will  be  on  the  front  dead 
center  A.  And  if  now  the  workmanship  of  the  valve  gear  is  absolutely  perfect,  the 
lead,  when  the  crank  is  in  this  position,  will  also  be  i^  of  an  inch.  But  to  obtain  such 
workmanship  is  very  difficult,  if  not  impossible,  and  therefore  it  must  notlje  surprising 
to  find  the  lead  at  the  front  end  of  the  cylinder  to  be  somewhat  incorrect.  Assume  the 
lead  at  the  front  end  is  £  of  an  inch,  and  at  the  back  end  n,  of  an  inch.  Now,  since, 
according  to  our  example,  n,  of  an  inch  is  the  required  lead,  and  since  the  lead  at  both 
ends  of  the  cylinder  must  be  equal,  it  follows  that  the  lead  at  the  front  end  must  be 
reduced,  without  changing  the  lead  at  the  back  end  of  the  cylinder.  To  do  this  we 
must  shorten  the  forward  eccentric-rod  ^  of  an  inch,  and  move  the  center  x  of  the 
eccentric  towards  the  center  line  B  C  of  the  crank,  until  the  valve  has  just  opened  the 
front  steam  port  rg  of  an  inch.  Turn  the  driving  wheel  once  or  twice  around  its  axis 
and  examine  the  lead ;  if  the  lead  is  equal  to  n,  of  an  inch  at  each  end  of  the  cylinder) 
consider,  for  the  present,  the  setting  of  the  forward  geaj*  to  be  correct.  Next  move  the 
reverse  lever  //  into  full  backward  gear,  and  find  the  correct  position  of  the  backward 
eccentric  y  in  a  similar  manner  to  that  adopted  for  finding  the  position  of  the  forward 
eccentric  x.  In  finding  the  position  of  the  backward  eccentric  y  or  setting  the  back- 
ward motion  of  the  valve  gear,  the  driving  wheel  should  always  be  turned  in  a  direction 
opposite  to  that  indicated  by  arrow  1.  Now  put  the  valve  motion  in  full-gear  forward, 
and  see  that  the  adjustment  of  the  backward  gear  has  not  deranged  the  valve  motion 
when  in  the  forward  gear.  Perhaps  a  little  readjustment  will  be  necessary ;  if  so,  the 
valve  motion  will  generally  indicate  where  the  readjustment  is  to  be  made.  When  the 
valve  motion  is  in  the  forward  gear,  the  driving  wheel  should  always  be  turned  in  the 
direction  of  the  arrow  1 — never  in  the  opposite  direction.  Should  the  wheel  be  turned 
a  little  further  than  was  intended,  we  must  either  turn  the  wheel  completely  around 
its  axis,  or  turn  it  back  considerably  beyond  the  stopping  point,  and  then  turn  slowly 
and  carefully  in  the  direction  of  the  arrow  1,  until  it  arrives  at  the  stopping  point. 
By  so  doing,  any  slack  motion  which  may  exist  will  not  interfere  with  the  correct 


MODERN  LOCOMOT1TK   CONSTRUCTION. 


133 


set ti iiir  <>t'  the  valve  motion.  If  now  it  is  foimd 
that  the  lead  is  ,-„  of  an  inch  at  each  end  of  the 
cylinder  when  the  valve  motion  is  in  full  back- 
ward gear,  and  also  in  full  forward  gear,  the  set- 
ting of  the  valve  gear  is  correct  and  complete. 
Since  we  have  assumed  that  the  position  of  the 
saddle-pin  <>M  the  saddle  is  correct,  and  that  the 
location  of  the  lifting  shaft  and  reverse  lever  is 
also  correct,  according  to  drawing  received  from 
the  drawing-room,  the  slide-valve  will  cut  off 
equal  portions  of  steam  at  each  end  of  the  cyl- 
inder. 

TO  LAY  OFF  THE  NOTCHES  IN  THE  QUADRANTS. 

154.  Iii  Fig.  204  we  have  represented  the 
same  mechanism  as  that  shown  in  Fig.  201,  but 
for  the  sake  of  clearness  some  of  the  details  have 
been  left  out. 

In  Fig.  204,  the  line  k  on  the  rear  end  of  the 
slides  represents  the  extreme  end  of  the  travel 
of  the  crosshead  end  jt  and  the  line  o  on  the  front 
end  of  the  slides  represents  the  extreme  end  of 
the  travel  of  the  crosshead  end  n.     It  is  now  to 
be  shown  how  the  notches  on  the  quadrants  G  can 
be  located,  so  that  when  the  reverse  lever  latch 
L  is  placed  in  one  of  these  notches 
steam  will  be  cut  off  when  the  cross- 
head  is  in  a  corresponding  given  posi- 
tion between  the  lines  k  and  o  marked 
on  the  slides. 

For  this  purpose  we  must  attach 
the  conneefcmg-rod  to  the  crank-pin 
and  crosshead,  so  that  when  the  wheel 
is  in  motion  the  crosshead  will  also  be 
in  motion. 

In  Art.  130  we  have  seen  that  there 
are  two  ways  of  arranging  the  notches 
in  the  quadrants  (I,  namely: 

1st.  The  notches  may  be  cut      •= 
as  close  together  as  the  strength 
of  material  will  allow. 

2d.  The  notches  may  be 
cut  so  as  to  hold  the  reverse 
lever  in  positions  which  will 
cause  the  steam  to  be  cut  off 


134  MODERN  LOCOMOTIVE   CONSTRUCTION. 

in  the  cylinder  at  some  full  number  of  inches  in  the  stroke.  We  will  first  consider 
how  to  locate  the  notches  in  the  quadrants  when  steam  is  to  be  cut  off  at  a  number 
of  full  inches  in  the  stroke. 

Assume  that  the  stroke,  in  Fig.  204,  is  24  inches,  and  that  steam  is  to  be  cut  off 
when  the  piston  has  traveled  6,  8, 10, 12,  and  15  inches  from  the  beginning  of  the  stroke ; 
it  is  required  to  find  the  positions  of  the  notches  in  the  quadrants  G,  and  also,  as  is 
always  customary,  find  the  center  notch  which  will  hold  the  valve  motion  in  mid-gear. 

In  Art.  152  we  have  shown  how  to  locate  the  notches  for  full-gear  forward,  and 
also  for  full-gear  backward ;  it  now  remains  to  be  shown  how  to  locate  the  intermediate 
notches.  First,  let  us  find  the  location  of  the  center  notch,  or,  in  other  words,  the 
notch  which  will  hold  the  valve  motion  in  mid-gear.  To  do  this  move  the  reverse 
lever  H  (Fig.  204)  towards  the  middle  of  the  quadrants  G,  and  in  the  meantime  turn 
the  driving  wheel  in  the  direction  of  arrow  1 — that  is,  in  the  direction  of  the  for- 
ward motion  of  the  wheel ;  also  watch  the  movement  of  the  slide-valve.  When  finally 
the  reverse  lever  H  arrives  in  a  position  in  the  central  portion  of  the  quadrants  which 
gives  the  shortest  travel  to  the  slide-valve,  the  reverse  lever  is  then  in  the  desired 
position,  and  will  hold  the  whole  valve  motion  in  mid-gear,  and  the  latch  of  the 
reverse  lever  will  indicate  the  position  of  the  center  notch ;  therefore  on  the  arcs  G 
scribe  off  the  end  of  the  latch  L,  and  to  these  marks  the  center  notch  must  be 
cut.  To  locate  the  other  notches  lay  off  on  the  slides  a  number  of  fine  and  dis- 
tinct lines,  as  represented  in  the  figure  by  the  lines  marked  6,  8,  etc.;  in  scribing 
these  lines  on  the  slides  care  should  be  taken  not  to  disfigure  them.  The  first  line, 
marked  6  in  our  figure,  must  be  6  inches  from  the  line  k ;  the  second  line,  marked  8, 
must  be  8  inches  from  the  line  k;  and  the  third  10  inches  from  A';  and  so  on  for  all 
the  given  points  of  cut-off.  Now  turn  the  driving  wheel  in  the  direction  of  the  arrow 
1,  causing  the  crosshead  to  move  away  from  k ;  as  soon  as  the  edge  j  of  the  crosshead 
touches  the  line  marked  6,  stop  the  motion  of  the  wheel,  and  then  slowly  move  the 
reverse  lever  H  from  the  center  notch  towards  the  forward  end  of  the  quadrants  G 
until  the  slide-valve  has  just  closed  the  rear  steam  port ;  then  on  the  arcs  mark  off  the 
end  of  the  latch,  and  thus  locate  the  notch  which  will  hold  the  valve  motion  in  a 
position  that  will  cut  off  steam  at  6  inches  from  the  beginning  of  the  stroke.  Again, 
turn  the  driving  wheel  in  the  direction  of  the  arrow  1,  and  when  the  edge  j  of  the 
crosshead  touches  the  line  marked  8  on  the  slide,  stop  the  motion  of  the  wheel,  and 
then  move  the  reverse  lever  towards  the  forward  end  of  the  quadrants  until  the  slide- 
valve  has  just  closed  the  rear  steam  port ;  then  again,  on  the  quadrants,  mark  off  the 
end  of  the  latch  L  and  thus  locate  the  8-inch  notch.  In  a  similar  manner  the  locations 
of  all  the  other  notches  in  the  forward  gear  are  obtained.  In  this  example  we  have 
assumed  that  the  valve  motion  has  been  correctly  designed  in  the  drawing-room, 
and  also  correctly  put  up  in  the  erecting  shop,  and  consequently  we  do  expect  that, 
when  the  driving  wheel  moves  in  the  direction  of  arrow  1,  and  the  crosshead  in  the 
direction  of  the  arrow  2,  and  the  reverse-lever  latch  is  in  the  notch  marked  6,  the 
valve  will  close  the  front  steam  port  when  the  crosshead  edge  n  has  moved  6  inches 
away  from  the  line  o,  which  was  previously  marked  on  the  slide ;  or  when  the  latch 
is  in  the  notch  marked  8,  the  valve  will  close  the  front  steam  port  when  the  crosshead 
edge  n  has  moved  8  inches  from  the  line  o. 


MODERN  LOCOMOTIVE   CONSTRUCTION.  135 

To  lay  off  the  notches  for  the  backward  motion  a  similar  method  is  pursued,  and 
the  same  lines  indicating  the  points  of  cut-off  previously  described  on  the  slides,  and 
used  for  the  forward  motion  of  the  valve  gear,  can  also  now  be  used  for  the  backward 
motion.  Hence,  to  lay  off  the  notches  for  the  backward  motion,  again  place  the 
reverse  lever  //  in  the  center  notch,  and  place  the  wheel  in  a  position  in  which  the 
edge  j  of  the  crosshead  will  touch,  or  nearly  so,  the  line  k  on  the  slides.  Then  turn 
the  driving  wheel  in  a  direction  opposite  to  that  of  arrow  1,  and  when,  the  edge  j  of 
the  crosshead  touches  the  h'ne  marked  6  on  the  slide,  stop  the  motion  of  the  wheel, 
and  move  the  reverse  lever  H  towards  the  back  end  of  the  quadrants,  until  the  slide- 
valve  closes  the  rear  steam  port ;  then  mark  off  on  the  quadrants  the  end  of  the  latch, 
thus  obtaining  the  location  of  the  6-inch  notch  for  the  backward  motion.  In  a 
similar  manner  the  position  of  all  the  notches  in  the  backward  gear  are  obtained; 
and  since  we  have  assumed  that  the  valve  gear  has  been  correctly  designed,  it  is 
expected  that  when  the  reverse  lever  latch  is  in  the  6,  8,  or  10  inch,  etc.,  notches 
steam  will  also  be  cut  off  at  6,  8,  or  10  inches  when  the  crosshead  travels  in  the 
opposite  direction,  namely,  in  the  direction  of  arrow  2.  Of  course,  a  perfect,  equal 
cut-off  is  hardly  to  be  expected ;  that  is,  when  the  crosshead  moves  in  the  direction  of 
arrow  3,  and  steam  is  cut  off  at  6  inches,  it  is  not  to  be  expected  that  steam  will  also 
be  cut  off  at  6  inches  when  the  crosshead  moves  in  the  direction  of  arrow  2 ;  but  we 
do  expect  that,  as  long  as  the  reverse-lever  latch  remains  in  one  notch,  the  points  of 
cut-off  will  not  vary  more  than  J  of  an  inch.  If  the  difference  is  greater  than  J  of  an 
inch,  an  inaccuracy  exists  in  the  fitting  up  of  the  valve  gear,  or  it  is  due  to  a  faulty 
construction ;  in  either  case  it  should  be  rectified. 

The  notches  must  be  correctly  located  for  both  cylinders,  so  that  when  steam  is 
cut  off  at  6,  10  inches,  etc.,  in  one  cylinder  the  steam  must  be  cut  off  at  the  same 
points  in  the  other  cylinder,  or,  in  other  words,  both  cylinders  must  receive  the  same 
amount  of  steam.  Hence,  after  the  notches  have  been  correctly  located  for  one  cylin- 
der, we  must  test  their  accuracy  for  the  other  cylinder;  and  if  then  one  cylinder 
receives  more  steam  than  the  other,  we  generally  find  that  the  cause  for  this  inac- 
curacy is  in  the  position  of  the  lifting-shaft  arms,  which  are  out  of  line — that  is,  one  of 
the  arms  may  hold  the  link  too  high  or  too  low,  and  consequently  this  arm  must  be 
sprung  into  the  proper  position,  which  is  often  accomplished  while  the  shaft  is  cold, 
without  taking  it  out  of  the  engine. 

When  steam  is  to  be  cut  off  equally  without  regard  to  any  particular  points  of 
cut-off  or  full  number  of  inches  in  the  stroke,  the  notches  are  spaced  off  evenly  be- 
tween the  two  outside  ones  and  as  close  together  as  the  strength  of  the  material 
will  allow,  and  therefore  generally  have  £  an  inch  pitch — that  is,  J  an  inch  from 
center  to  center  of  notch.  (See  Art.  130.)  In  this  case  also  the  same  care  must  be 
taken  to  arrange  the  valve  motion  in  a  manner  which  will  supply  both  cylinders  with 
tlie  same  amount  of  steam. 

It  should  be  remembered  that  the  foregoing  instructions  in  setting  the  valve  gear 
are  intended  for  new  work  only.  When  the  valve  gear  for  an  old  engine  is  to  be 
readjusted,  the  wear  of  the  different  parts  must  lie  taken  into  consideration,  which 
sometimes  makes  the  readjustment  a  difficult  process.  re<|uiring  experience  and  a 
knowledge  of  the  principles  connected  with  the  design  of  the  valve  motion. 


136 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


REMAKES   KELATING  TO    THE  VALVE   MOTION    IN   WHICH    THE    ECCENTRICS    ARE    NOT    ON   THE 

MAIN   DRIVING  AXLE. 

155.  Sometimes  it  happens  that  in  designing  small  locomotives  the  eccentrics 
cannot  be  placed  on  the  main  axle,  consequently  we  occasionally  meet  with  a  valve 
gear  as  shown  in  Fig.  205.  In  this  valve  gear  the  eccentrics  are  placed  on  the  axle 
in  front  of  the  main  axle.  Placing  the  eccentrics  on  the  front  axle  instead  of  the 
main  axle  will  not  change  the  previous  rules  we  have  given  for  designing  the  valve 
motion;  but  it  should  be  noticed  that,  in  order  to  preserve  the  lead,  no  matter  in 
what  position  the  link  may  be  placed,  as  explained  in  Art.  151,  the  eccentric-rods  are 


connected  to  the  link  in  a  manner  which  will  put  the  lower  end  of  the  link  in  action 
when  the  engine  is  going  ahead,  or,  in  other  words,  the  link  must  be  raised  when  the 
engine  is  to  run  forward,  and  lowered  when  the  engine  is  to  run  backward.  Of  course 
the  eccentric-rods  could  be  connected  to  the  link  so  that  it  would  be  in  the  lower 
position  when  going  ahead,  and  in  the  upper  position  when  running  backward — that  is, 
the  link  could  be  made  to  occupy  a  similar  position  to  that  in  the  link  motions  previ- 
ously illustrated,  and  for  that  matter  would  also  work  all  right  while  in  full-gear,  but 
the  lead  will  be  destroyed  when  the  link  is  placed  in  any  of  the  intermediate  positions 
for  the  same  reasons  as  explained  in  Art.  151 ;  and  since  we  must  have  lead  in 
whatever  position  the  link  is  placed,  we  must  suspend  the  link  as  shown  in  Fig.  205 


MODERN  LOCOMOTIVE   COXSTRVCTIOX. 

It  will  also  be  found  that  in  locating  the  lifting  shaft  according  to  the  rules  previously 
given,  it  will  have  to  be  placed  in  front  of  the  link,  instead  of  being  placed  in  the 
ivur  of  it  as  has  been  found  necessary  in  all  the  foregoing  link  motions.  But  by 
placing  the  lifting  shaft,  in  the  case  now  under  consideration,  in  front  of  the  link,  we 
will  gain  an  additional  advantage,  namely,  in  connecting  the  lifting-shaft  arm  to  the 
reverse  lever,  the  latter  will  be  made  to  point  ahead  when  the  engine  is  going  forward 
— an  important  fact  which  should  not  be  neglected  or  forgotten,  because  if  the  valve 
motion  should  happen  to  be  arranged  in  a  manner  which  would  cause  the  reverse  lever 
to  point  back  while  the  engine  is  going  ahead,  the  engineer  may  be  thrown  into  con- 
fusion when,  in  an  emergency,  the  engine  has  to  be  quickly  reversed,  and  thus  lead 
to  accidents. 


CHAPTER    IV. 


PISTONS.— -CROSSHEADS.— SLIDES.— STUFFING  BOXES. 


PISTONS. 


156.  Locomotive  pistons  consist  of  two  principal  parts,  namely,  the  piston  head 
and  the  packing.  Locomotive  pistons  may  also  be  divided  into  three  classes,  accord- 
ing to  the  construction  of  the  head;  the  first  class  embracing  all  those  in  which 
the  piston  head  is  made  up  of  more  than  one  piece,  such  as  shown  in  Figs.  206  and 


BED 


Fiff.20(i 


Fig.20? 


207 ;  this  class  of  pistons  we  may  call  the  "  built-up  piston."  The  second  class  will 
embrace  the  pistons  in  which  the  head  is  cast  hollow,  such  as  shown  in  Fig.  208 ;  this 
class  of  piston  we  may  call  the  hollow  piston.  To  the  third  class  belong  the  solid 
piston,  such  as  is  shown  in  Fig.  209. 

THE  BUILT-UP  PISTON. 

157.  Figs.  206  and  207.  In  this  figure  the  piston  head  is  made  up  of  the  following 
pieces :  The  spider,  which  is  marked  A ;  the  follower  plate,  or  simply  the  follower, 
marked  B ;  and  the  follower  bolts  C,  which  unite  the  follower  to  the  spider. 

In  American  locomotives  the  spider  A  is  made  of  cast-iron.  It  should  be  made, 
in  fact  the  whole  piston  should  be  made,  as  light  as  possible,  and  consequently  its 
metal  must  be  judiciously  distributed,  so  as  to  obtain  the  required  strength  with  the 
least  amount  of  metal.  By  so  doing,  less  counterbalance  will  be  required,  besides 


MODEK\  LOCOMOTirK   COXSTRUCTION. 


139 


gaining  other  advantages.  The  spider  is  generally  keyed  to  the  piston-rod  by  the  key 
/;  sometimes  the  piston  is  secured  to  the  rod  by  means  of  a  nut,  as  shown  in  Fig.  208. 
In  very  large  pistons,  such  as  are  used  in  stationary  or  marine  engines,  the  depth  of 
the  spider  (corresponding  to  the  depth  x  in  Fig.  206)  is  computed  for  strength ;  the 
width  of  the  packing  (corresponding  to  the  width  y  in  Fig.  206) — that  is,  the  distance 
from  the  face  of  the  flange  of  the  spider  to  the  face  of  the  follower — is  generally  made 
equal  to  £.  to  ^  of  the  diameter  of  the  piston ;  the  latter  proportion  being  adopted  for 
pistons  of  comparatively  small  diameter  (as  large  as  a  locomotive  piston  22  inches  in 
diameter)  and  the  former  proportion  for  pistons  of  very  large  diameter.  Consequently, 
in  stationary  engines,  and  also  in  marine  engines,  we  often  find  pistons  whose  depth 


FiO.%08 


Vifl.%09 


at  the  center  is  greater  than  that  at  the  circumference,  so  that  when  one  face  of  these 
pistons  is  flat  the  other  face  will  be  part  of  a  spherical  surface.  In  locomotive 
pistons  the  depth  x  of  the  spider  is  determined  by  the  key  which  holds  it  to 
the  piston-rod.  The  dimensions  of  these  keys,  depth  of  spider,  and  size  of  piston- 
rods  are  given  in  Figs.  220  to  232.  Now,  practically,  in  locomotive  pistons  of  small 
diameter  the  necessary  depth  x  of  spider  does  not  give  us  an  excessive  depth  for  the 
packing,  and  in  those  of  larger  diameter  the  depth  x  will  give  a  depth  of  packing 
agreeing  with  the  proportions  given  for  stationary  and  marine  engines.  Therefore 
the  faces  s  and  s.>  of  locomotive  pistons  are  generally  flat,  sometimes  slightly 
tapered  towards  the  circumference,  as  shown  in  the  figures. 

The  follower  plates  (Figs.  206  and  207)  are  made  of  cast-iron,  and  usually  cover 
the  whole  open  end  of  the  spider  when  follower  bolts  are  used,  as  shown  in  the  figure. 

FOLLOWER  BOLTS. 


158.  In  any  mechanism  having  such  a  lii^li  rate  of   reciprocating  motion  as   a 
locomotive  piston,  bolts,  unless  carefully  fitted,  are  liable  to  become  loose  and  work 


140  MODERN  LOCOMOTIVE   CONSTRUCTION. 

out.  To  prevent  the  follower  bolts  from  becoming  loose,  they  are  extended  into  the 
spider  through  a  distance  which  is  at  least  equal  to  three  times  the  diameter  of  the 
bolt,  and  great  care  is  taken  to  make  a  good  fit  of  the  thread.  Sometimes,  on  account 
of  the  rust,  trouble  is  experienced  when  the  follower  bolts  are  to  be  taken  out  of  the 
spider.  To  obviate  this  difficulty  a  few  locomotive  builders  insert  brass  plugs  into  the 
spider.  (See  Fig.  210.)  These  plugs  are  tightly  screwed  into  the  spider  and  then 
riveted  over  at  the  ends.  In  these  brass  plugs  holes  are  drilled  and  tapped  to  receive 
the  follower  bolts,  preventing  any  rust  from  causing  a  difficulty  in  removing  the  bolts 
when  it  becomes  necessary  to  do  so. 

The  construction  of  the  hollow  piston,  shown  in  Fig.  208,  needs  no  explanation, 
with  the  exception  to  state  that  a  number  of  core  holes  B  are  cast  in  one  side  of  the 
piston,  and  when  the  piston  is  cleaned  the  core  holes  are  tapped  and  plugged  up. 

The  solid  piston  shown  in  Fig.  209  is  used  when  extreme  lightness  is  required. 
This  form  of  piston  is  suitable  for  cast-iron,  wrought-irou,  or  steel. 

PISTON  PACKING. 

159.  A  great  amount  of  thought,  labor,  and  expense  has  been  expended  in  con- 
structing and  perfecting  a  piston  packing  which  shall  give  satisfaction  under  all 
circumstances,  and  consequently  many  different  kinds  of  piston  packing  are  in  use. 
But  we  have  noticed  that  the  kind  of  piston  packing  which  has  given  the  best  satisfac- 
tion so  far  is  one  in  which  two  distinguishing  features  are  found,  namely,  durability 
and  simplicity ;  and  indeed  these  two  features  we  should  expect  to  find  in  any  good 
piston  packing,  because  any  packing  that  will  not  remain  steam  tight  for  a  consider- 
able length  of  time  is  troublesome,  annoying,  and  unprofitable ;  and  a  packing  which 
does  not  possess  the  feature  of  simplicity  is  dangerous,  particularly  so  in  fast-running 
engines.  The  packing  in  the  hollow  piston,  Fig.  208,  consists  simply  of  two  cast-iron 
rings  D  7),  which  are  sprung  into  the  grooves  turned  into  the  periphery  of  the  piston. 
Each  ring  is  cut  in  one  place ;  the  favorite  form  of  the  cut  is  shown  in  Figs.  214  and 
215.  This  piston,  on  account  of  its  great  simplicity,  is  a  favorite  one  on  many  roads. 
It  is  more  suitable  for  cylinders  of  large  diameter  than  smaller  ones,  because  in  the 
latter  there  is  danger  of  breaking  the  rings  in  attempting  to  spring  them  into  the 
grooves.  Similar  packing  rings  are  used  for  the  solid  piston  shown  in  Fig.  209. 

In  Figs.  206  and  207  we  have  represented  another  piston  packing  which,  on 
account  of  its  simplicity  and  durability,  has  in  late  years  gained  great  favor  among 
engineers,  and  is  certainly  an  excellent  piston  packing  for  locomotives.  It  consists 
simply  of  two  cast-iron  rings  D  D,  and  one  cast-iron  T-ring  marked  E.  We  call 
this  ring  a  "  T-ring  "  because  its  cross-section  resembles  a  T.  The  rings  D  D  are  cut 
open  in  one  place,  and  consequently  are  called  single-cut  rings.  Two  ways  of  cutting 
these  rings  are  generally  adopted.  Some  master-mechanics  will  have  the  rings  D  D 
cut  as  shown  in  Fig.  213.  Here  a  hole  F  F  is  first  drilled  and  made  oblong  in  the 
direction  of  F  F,  then  through  the  center  of  the  hole  the  ring  is  cut  open  square 
across.  Care  should  be  taken  to  cut  out  a  sufficient  amount  of  metal,  so  that  when 
the  packing  ring  is  placed  in  the  cylinder  the  two  ends  will  not  touch  each  other, 
but  will  leave  an  opening  sufficiently  great  to  allow  for  the  expansion  of  the  ring 


MOI>I-:I;\  tOCOJtOJtrK  ro.v,sr/?rcrro,v. 


141 


when  it  becomes  liot.  The  same  remark  applies  to  the  hole  F  F;  this  should  be 
made  long  enough,  as  shown,  to  allow  a  sufficient  clearance  around  the  pin  which 
is  inserted  in  this  hole  and  driven  into  the  T-ring,  otherwise  the  expansion  will  force 
the  ends  of  the  packing  ring  against  the  pin  and  prevent  the  ring  from  acting  as  it 
should  do.  The  object  of  this  pin  is  to  prevent  the  ring  Z)  from  moving  around  the 


Fig.210 


Fig.2'11 


Fig.Zlit 


Fig.218 

T-ring.  When  the  rings  D  T)  are  cut  in  the  manner  just  described,  they  are  turned 
about  i\  to  ,r'(1  of  an  inch  larger  in  diameter  than  the  bore  of  the  cylinder,  so  as  to 
give  the  rings  an  inherent  elasticity.  The  other  way  of  cutting  these  rings  is  shown 
in  Fig.  214.  Here  half  the  width  of  the  ring  is  cut  out  through  a  distance  of  about  5 
of  an  inch,  which  enables  the  ends  of  the  ring  to  overlap  each  other  about  f  of  an 
inch.  The  writer  believes  (hut  when  a  ring  is  to  be-  cut  in  this  manner  the  best  method 
to  pursue  will  be  as  follows:  First  turn  the  ring  large  enough  in  diameter  to  allow  for 


142  MODERN  LOCOMOTIVE   CONSTRUCTION. 

the  metal  to  be  cut  out.  Then  cut  the  openings  A  and  B,  Fig.  215,  through  half  the 
width  of  the  ring,  and  each  about  J  of  an  inch  long.  When  these  are  cut  press  the 
ends  together,  so  that  the  ends  will  overlap  each  other  about  f  of  an  inch ;  through 
these  reduced  ends  drive  a  small  temporary  pin,  and  then  again  turn  the  ring  to  fit 
the  bore  of  the  cylinder ;  after  that  take  out  this  pin,  as  it  has  served  its  purpose. 
Other  pins  are  driven  in  the  T-ring,  which  fit  loosely  in  holes  drilled  in  the  packing 
rings,  to  prevent  the  latter  from  turning  around  the  T-ring. 

As  to  the  proportions  of  the  cross-section  of  the  rings  D  D  (see  Fig.  206),  a  differ- 
ence of  opinion  exists.  Some  master-mechanics  make  the  width  of  the  ring  greater 
than  the  depth,  others  will  insist  on  having  the  width  smaller  than  the  depth,  but 
more  frequently  we  find  the  cross-section  to  be  a  square.  The  writer  believes  that 
packing  rings  whose  cross-section  is  £  inch  square  will  give  good  satisfaction  for 
cylinders  9  inches  in  diameter ;  and  this  cross-section  should  be  gradually  increased 
for  cylinders  of  larger  diameter,  in  a  manner  which  will  bring  the  cross-section  of  the 
ring  to  f  inch  square  for  locomotive  cylinders  22  inches  in  diameter.  In  the 
majority  of  locomotives  the  depth  H  of  lings  D  D  remains  the  same  throughout,  as 
shown  in  Fig.  216.  They  are  made  to  press  against  the  interior  surface  of  the 
cylinder  by  their  inherent  elasticity,  and  when  this  elasticity  is  exhausted  through 
constant  wear,  it  can  be  partially  restored  by  a  few  judicious  light  blows  of  the 
hammer  in  the  inside  of  the  ring.  Since  these  rings  will  wear  most  at  the 
opening,  and  in  time  wear  the  cylinders  oval  instead  of  cylindrical,  it  has  been 
assumed  that  they  have  not  a  uniform  pressure  all  around  against  the  cylinder 
surface,  and  this  is  true.  Hence,  on  account  of  this  unequal  pressure,  some  master- 
mechanics  make  these  packing  rings  of  the  form  as  shown  in  Fig.  217,  in  which  the  depth 
H  at  the  opening  of  the  ring  is  less  than  the  depth  H  opposite  the  opening.  The  object 
aimed  at  is,  of  course,  to  obtain  as  nearly  as  possible  an  equal  pressure  of  the  ring 
against  the  interior  surface  of  the  cylinder.  But  from  the  writer's  experience,  and 
upon  careful  inquiry,  he  is  led  to  believe  that  rings  of  uniform  depth  throughout,  as 
shown  in  Fig.  216,  will  give  as  good  satisfaction  as  the  ring  shown  in  Fig.  217. 

Cast-iron  is  generally  considered  to  be  as  good  a  metal  as  can  be  used  for  piston 
packing  rings,  as  it  wears  well,  particularly  when  the  metal  in  the  cylinder  is  hard, 
as  it  should  be;  furthermore,  rings  of  the  kind  just  described  need  but  very  little 
inspection,  and  can  be  used  for  a  considerable  length  of  time. 

160.  Sometimes  the  packing  rings  are  cut  into  four  or  more  pieces,  and  made  to 
press  against  the  interior  cylinder  surface  by  the  aid  of  springs;  the  advantage 
claimed  for  this  arrangement  is  an  equal  pressure  against  the  cylinder  surface. 

/ 

DUNBAR  PACKING. 

A  notable  packing  of  this  kind  is  the  Dunbar  packing  shown  in  Fig.  219A.  The 
packing  consists  of  two  kinds  of  rings:  one  has  an  L-shaped  cross-section  and  the 
other  a  square  section.  Each  ring  is  cut  into  a  number  of  sections — in  this  case  six  in 
number — and  when  put  together  the  joints  are  staggered.  These  sections  are  pressed 
out  by  a  number  of  springs  made  of  round  steel  wire  extending  around  the  T-ring ;  the 
latter  is  of  the  same  form  as  is  usually  used  for  the  ordinary  class  of  pistons.  But  cut- 


XODERX  LOCOMOTIVE   COSSTRUCTION. 


143 


Fig.  219  A 


Platan  for  10  In.  Cylinder 
"Dunbar" 


ting  the  rings  into  so  many 
pieces,  and  then  introduc- 
ing springs,  interferes  with 
the  simplicity  of  construc- 
tion, and  therefore  pack- 
ings of  this  kind  have  not 
been  as  frequently  used  in 
late  years  as  formerly. 

T-RING. — POSITION  OF  PACK- 
ING KINGS. 

1  til.  Fig.  206.  The  outer 
smi'ace  of  the  T-ring  is 
turned  to  fit  easily  the  bore 
of  the  cylinder,  so  that 
when  the  T-ring  becomes 
hot  its  expansion  will  not 
interfere  with  its  free  motion  in  the  cylinder.  The  width  of  this  ring  must  be 
turned  to  such  dimensions  that  it  can  be  held  firmly  in  position  by  the  follower  plate, 
when  the  latter  is  screwed  to  the  spider.  The  recesses  turned  at  each  corner  of  the  ring 
will  form  the  grooves  for  the  packing  rings  when  the  piston  is  put  together.  Of  course, 
the  T-ring  is  not  cut  open.  The  advantage  obtained  by  the  use  of  this  ring  is  that  the 
packing  rings  can  be  readily  placed  in  or  taken  out  of  the  piston,  without  forcing  the 
ends  of  the  packing  rings  apart  to  place  them  into  the  grooves,  as  must  be  done  when 
the  solid  piston  is  used.  The  packing  rings  must  fit  in  the  grooves  very  accurately, 
not  too  loose,  yet  free  enough  to  allow  the  rings  to  adjust  themselves  to  the  interior 
surface  of  the  cylinder.  When  both  the  packing  rings  are  placed  in  position,  their 
openings  or  cuts  should  not  be  in  a  straight  line,  but  should  be  staggered,  having  a 
distance  of  5  or  6  inches  between  them,  as  indicated  by  distance  between  the  openings 
/  /•'  in  Fig.  206.  It  is  always  best  to  allow  the  openings  of  the  rings  to  lie  in  the 
lower  part  of  the  cylinder ;  but  in  no  case  should  they  be  placed  over  a  steam  way. 
Tin-  reason  for  this  is  that,  since  the  rings  are  turned  a  little  larger  in  diameter 
than  the  bore  of  the  cylinder,  they  must  be  compressed  in  placing  them  into  the 
cylinder ;  and  now,  should  the  hold  or  grip  upon  them  be  lost  through  some  accident 
or  mishap  while  in  the  act  of  placing  them  in  position,  the  ends  of  the  rings  may  fly 
into  the  steam  way,  and  generally  in  such  cases  they  must  be  ruined  by  cutting  them 
into  several  pieces,  in  order  to  take  the  rings  out. 


BRASS   PISTON  PACKING. 

162.  Another  very  good  packing  for  locomotive  pistons  is  shown  in  Figs.  211  and 
212.  In  former  years  this  packing  was  the  favorite,  and  even  in  late  years  some 
master-mechanics  show  a  reluctance  in  giving  it  up,  and  consequently  it  is  at  present 
frequently  met  with.  This  packing  consists  of  two  brass  packing  rings  /'  /',  one  inside 
cast-iron  ring  M,  the  packing  bolts  K  with  nuts,  and  the  packing  springs  L.  The 


144  -MODERN  LOCOMOTIVE   CONSTRUCTION. 

packing  rings  P  P  have  grooves  G  G  turned  in  them,  and  these  grooves  are  filled 
with  Babbitt  metal.  This  metal  will  prevent  the  scratching  of  the  interior  of  the 
cylinder  surface,  such  as  would  be  the  case  when  brass  alone  is  used.  The  rings  P  P 
are  turned  all  over  very  accurately ;  the  width  of  each  of  these  rings  is  equal  to  half 
the  space  between  the  flange  of  the  spider  and  the  follower  plate,  so  that  the  rings 
will  fill  this  space  without  binding,  and  have  freedom  to  adjust  themselves  to  the 
bore  of  the  cylinder  as  the  piston  moves  forward  and  backward.  Consequently — 
as  will  be  seen  by  the  illustrations,  Figs.  220  to  232 — the  width  of  each  packing  ring 
for  large  pistons  will  be  l£  inches,  and  for  the  smaller  pistons  In,  inches.  The  depth 
of  these  rings  is  generally  equal  to  f  of  their,  width.  Each  ring  P  P  is  cut  open  in 
one  place,  usually  at  an  angle  as  shown  at  F  F,  Fig.  211.  The  diameter  of  these 
packing  rings  should  be  sufficiently  large  to  allow  only  for  the  amount  of  metal  that 
is  to  be  cut  out  of  them,  so  that  when  the  ring  is  placed  in  the  cylinder  the  ends 
of  the  opening  will  be  sufficiently  apart  to  prevent  the  expansion  from  forcing  the 
ends  together. 

The  cast-iron  ring  M  is  turned  to  fit  the  inside  of  the  brass  rings  P  P,  is  cut  in 
one  place  square  across,  and  its  width  must  be  exactly  equal  to  the  sum  of  the  widths 
of  the  brass  rings,  so  as  to  give  it  freedom  to  adjust  itself  to  the  inside  of  the 
brass  rings.  The  thickness  of  the  ring  M  is  usually  about  %  of  an  inch  for  large 
pistons,  and  about  f%  for  smaller  ones.  The  purpose  of  the  ring  M  is  to  furnish 
a  bearing  for  the  springs  L  L,  and  to  distribute  their  pressure  equally  on  the  packing 
rings;  also  to  form  a  steam-tight  joint  with  the  interior  of  the  brass  rings  P  P. 
Dowel  pins  N,  Fig.  212,  are  driven  into  the  ring  M  and  allowed  to  project  into 
holes  drilled  into  the  brass  rings  P  P,  preventing  the  latter  from  moving  around  the 
ring  M. 

The  brass  rings  P  P  have  but  very  little  inherent  elasticity,  and  consequently  the 
packing  springs  L  L  are  needed  for  the  purpose  of  pressing  the  packing  rings 
against  the  interior  cylinder  surface.  In  large  locomotive  pistons  the  springs  L  L 
are  usually  about  5  inches  long,  2j  inches  wide,  and  £  of  an  inch  thick  in  the 
center ;  the  thickness  is  reduced  towards  the  ends.  For  smaller  pistons  these  springs 
are  about  4  inches  long,  2  inches  wide,  and  J  of  an  inch  thick. 

The  packing  bolts  K  K  are  generally  f  of  an  inch  in  diameter ;  their  heads,  which 
are  of  the  T  form,  are  set  into  grooves  cast  into  the  hubs  of  the  spider,  which  will 
prevent  the  bolts  from  turning  when  the  nuts  are  screwed  against  the  springs  to 
press  out  and  adjust  the  packing  rings  as  may  be  required. 

PISTON-RODS. 

163.  Locomotive  piston-rods,  and,  in  fact,  piston-rods  for  nearly  all  kinds  of 
steam  engines,  are  subjected  alternately  to  a  tensile  and  compressive  stress.  By 
tensile  stress  is  meant  that  force  which  tends  to  produce  fracture  by  pulling  or  tearing 
the  piston-rod  apart ;  and  by  compressive  stress  is  meant  that  force  which  tends  to  pro- 
duce fracture  by  crushing  the  rod.  In  calculating  the  strength  of  any  piston-rod  the 
tensile  and  also  the  oompressive  stress  to  which  it  is  subjected  must  be  considered ; 
thus,  for  instance :  We  first  find  the  diameter  which  will  give  the  piston-rod  sufficient 


MODESX  LOCOMOTIVE   CONSTRUCTION.  145 

strength  to  resist  the  tensile  stress;  then  we  find  the  diameter  which  will  give  the 
piston-rod  sufficient  strength  to  resist  the  compressive  stress.  If  we  find  that  the 
diameter  must  be  3£  inches  to  resist  tensile  stress,  and  only  3  inches  to  resist  com- 
pressive stress,  then  the  diameter  of  the  rod  must  be  made  equal  to  3£  inches, 
otherwise  there  is  danger  of  producing  fracture  by  tearing;  if,  on  the  other  hand, 
we  find  that  the  diameter  of  the  piston-rod  must  be  3  inches  to  resist  compressive 
stress,  and  only  2J  inches  to  resist  tensile  stress,  then  the  diameter  of  the  piston-rod 
must  be  3  inches,  otherwise  there  is  danger  of  buckling  or  producing  fracture  by 
crushing  it.  From  these  remarks  we  may  conclude  that  in  some  cases  the  diameter 
of  the  piston-rod  is  determined  and  limited  by  the  tensile  stress,  and  in  other  cases 
by  the  compressive  stress.  Generally  speaking,  long  rods — that  is,  rods  which  are 
long  in  comparison  with  their  diameters — cannot  resist  as  much  compressive  stress 
as  tensile  stress,  consequently  their  diameters  are  determined  by  the  former;  short 
rods,  on  the  other  hand,  cannot  usually  resist  as  much  tensile  stress  as  compressive 
stn-ss,  hence  the  diameters  of  short  rods  are  determined  by  the  tensile  stress.  Loco- 
motive rods  are  comparatively  short  when  compared  with  their  diameters,  and  their 
ends  are  weakened  by  key  ways  or  threads ;  therefore  the  diameters  are  determined  "by 
the  tensile  stress  alone,  and  consequently  in  the  following,  compression  will  be  left  out 
of  consideration. 

In  speaking  of  the  diameter  of  a  piston-rod,  we  mean  the  diameter  of  that  part  of 
the  rod  which  reaches  from  the  piston  to  the  crosshead. 

164.  When  built-up  pistons  are  used,  the  piston  is  keyed  to  the  rod  in  the  majority 
of  cases ;  occasionally  we  find  the  two  united  by  a  nut.     The  ends  of  the  rod  which 
fit  in  the  piston  and  crosshead  are  made  smaller  in  diameter  than  that  of  the  rod,  so 
as  to  form   shoulders.     The   object  of  these   shoulders   is  twofold:    first,  they  will 
allow  the  rods  to  be  re-turned  when  that  becomes  necessary  through  constant  wear ; 
and  secondly,  it  is  considered  to  be  good  practice  to  have  a  shoulder  against  which 
the  piston  can  be  driven.     The  ends  are  tapered  f  of  an  inch  in  4  inches — that  is,  in  4 
indies  the  diameter  decreases  £  of  an  inch.     The  end  of  the  rod  which  fits  in  the 
piston  must  be  made  as  short  as  the  design  will  permit,  so  as  to  reduce  the  depth  of 
the  piston  as  much  as  possible.     The  crosshead  end  must  be  made  a  little  longer  than 
the  piston  end,  because  in  the  former  the  distance  between  the  key  and  the  shoulder  of 
the  piston-rod  must  be  increased  to  obtain  sufficient  metal  in  the  crosshead  hub  from 
the  key  to  the  end  of  the  hub.     The  taper  of  the  keys  is  \  of  an  inch  in  12  inches. 

165.  Figs.  220  to  232  inclusive  form  an  illustrated  table  of  the  pistons  and  rods 
from  which  all  the  principal  dimensions  can  at  once  be  obtained.     In  making  this 
table  the  writer  obtained  the  dimensions  of  rods  and  pistons  used  in  modern  locomo- 
tives and  doing  good  work,  but,  as  was  to  be  expected,  there  was  found  to  be  no 
uniformity  of  proportions;    but   yet  these  proportions  seemed  to  indicate  that  the 
tensile  stress  should  not  be  greater  than  10,000  pounds  per  square  inch  on  the  weakest 
part  of  the  piston-rod.     With  this  data  and  these  proportions  as  a  basis  the  writer  formed 
this  table,  in  which  the  dimensions  have  been  obtained  by  calculation,  so  that  when  these 
dimensions  are  adopted  and  the  pressure  in  the  cylinder  is  120  pounds,  the  tensile, 
stress  per  square  inch  on  the  weakest  part  of  the  piston-rod  and  key  will  not  exceed 
10,000  pounds. 


MHIIKRN  LOCOMOTITE    CONSTRUCTION.  147 

Mostly  all  the  dimensions  given  in  these  figures  agree,  and  the  others  very  nearly 
agree,  with  the  sizes  of  pistons  and  rods  used  in  first-class  locomotives.  The  writer 
believes  this  table  to  be  reliable. 

Locomotive  piston-rods  are  made  of  iron,  steel,  and  cold  rolled  iron.  When  cold 
rolled  iron  is  used  the  piston-rods  are  not  turned,  the  iron  being  rolled  to  the 
required  size. 

In  order  to  reduce  the  number  of  patterns  as  much  as  possible,  it  is  customary  to 
retain  the  same  diameter  for  steel  and  iron  piston-rods  for  a  given  diameter  of  cylinder ; 
that  is  to  say,  when  a  steel  rod  is  to  be  used  in  place  of  an  iron  rod,  the  diameter  of 
the  former  is  made  the  same  as  that  of  the  latter. 

Consequently,  our  remarks  in  regard  to  strength  of  locomotive  piston-rods  apply 
to  iron  as  well  as  steel  rods.  But  the  reader  must  not  be  led  to  understand  that  there 
is  no  difference  between  the  strength  of  steel  and  iron  rods,  and  that  in  designing 
piston-rods  for  other  engines  the  difference  between  the  strength  of  steel  and  iron  can 
be  neglected.  We  simply  wish  to  be  understood  that  in  locomotive  practice  only,  the 
difference  between  the  strength  of  steel  and  iron  rods  is  left  out  of  the  calculation,  so 
that  the  rod  made  of  the  weakest  material  will  still  be  strong  enough  to  do  the  work, 
and  thus  enable  us  to  establish  an  interchangeability  of  wider  range  of  the  mechanism 
whose  dimensions  depend  upon  the  diameter  of  the  piston-rod,  and  also  reduce  the 
number  of  patterns. 

It  is  the  general  practice  to  allow  on  the  piston-rod  a  tensile  stress  of  5,000 
pounds  per  square  inch,  and  not  to  exceed  this.  But  it  should  also  be  remembered 
that  new  locomotive  piston-rods  are  usually  made  £  of  an  inch  larger  in  diameter  than 
is  necessary  for  the  strength  of  the  rod,  so  that  in  case  the  rod  needs  to  be  turned 
down  on  account  of  wear,  it  will  still  be  strong  enough  to  do  the  work.  Conse- 
quently, after  having  found  the  correct  diameter  which  the  strength  of  the  rod 
demands,  we  must  increase  this  diameter  by  i  of  an  inch,  which  will  be  the  allowance 
for  wear.  Hence,  when  it  is  desirable  to  determine  by  calculation  the  diameter  of  a 
piston-rod  suitable  for  a  size  of  cylinder  and  a  steam  pressure  not  given  in  the  table, 
the  following  rule  may  be  employed : 

RULE  18. — Multiply  the  area  in  square  inches  of  the  piston  by  the  maximum  steam 
pressure  per  square  inch  in  the  cylinder;  the  product  will  be  the  total  pressure  on  the 
piston,  and  therefore  the  total  tensile  stress  on  the  piston-rod. 

Divide  this  product  by  5,000 ;  the  quotient  will  be  the  area  in  square  inches  of 
the  cross-section  of  the  piston-rod ;  the  corresponding  diameter  of  this  area,  and  an 
addition  of  J  of  an  inch  to  this  diameter,  will  be  required  diameter  of  the  piston-rod. 
Putting  this  rule  in  the  shape  of  a  formula,  we  have : 

Area  of  the  piston  in  square  inches  x  pressure  per  square  inch 

— =  area  of  the  piston-rod,  without  allowance  for  wear ; 

then, 

diam.  of  area  found  +  £  of  an  inch  =  required  diameter  of  piston-rod. 

EXAMPLE  44. — What  should  be  the  diameter  of  a  piston-rod  for  a  locomotive 
cylinder  18  inches  in  diameter ;  steam  pressure  in  cylinder,  120  pounds  per  square  inch? 

1254.47  x  120 

"000      ~  ~        ^  'area  °*  piston-rod,  without  allowance  tor  wear. 


148 


MODERN  LOCOMOTIVE   CONSTRICTION. 


Diameter  of  an  area  of  6.1  square  inches  =  2yf  inches,  nearly. 

2H  +  4  =  2-J-f  inches  =  required  diameter  of  piston-rod. 

In  Fig.  229  we  find  this  diameter  to  be  3  inches.  There  are  in  this  table  several 
diameters  which  will  slightly  exceed  the  diameters  found  by  this  rule.  The  reason  of 
this  is  that  in  making  this  table  the  writer  has  followed  the  general  practice  of  locomo- 
tive builders,  namely,  avoiding  -^  of  an  inch  in  the  diameters  of  the  piston-rods ;  also 
making  the  increase  of  the  diameters  of  rods  for  the  different  sizes  of  cylinders  as 
gradual  as  possible.  The  piston-rod  in  Fig.  232  is  about  £  of  an  inch  less  in  diameter 
than  would  be  obtained  by  calculation. 

We  have  given  this  small  diameter  because  such  was  used  in  the  few  engines  of 
this  size  that  we  have  met.  The  figures  obtained  by  the  rule  will  give  sufficient 
strength  to  the  piston-rods ;  the  figures  in  table  agree  closer  with  practice.  It  will 
be  well  to  remark  here  that  the  area  of  the  weakest  part  of  the 
piston-rod  is  practically  equal  to  one-half  the  area  of  the  cross- 
section  of  the  rod;  and  this  remark  applies  equally  well  to  the 
piston-rods  which  are  keyed  to  the  piston,  and  those  which  are 
united  to  it  by  a  nut.  Thus,  for  instance:  Let  Fig.  233  repre- 
sent one  end  of  the  piston-rod ;  then  the  section  through  a  b  will 
obviously  be  the  weakest  part  of  the  rod  to  resist  tensile  stress ; 
and  it  will  generally  be  found  that  the  area  of  this  section  is  equal 
to  one-half  the  area  of  the  section  through  c  d.  Or,  again,  when 
the  piston-rod  is  united  to  the  piston  by  a  nut  (such  a  rod  is  repre- 
sented in  Fig.  219),  then  the  area  of  a  section  through  the  bottom 
of  the  thread,  which  is  the  weakest  part,  will  be  equal  to  about  one- 
half  the  area  of  the  rod.  Consequently  it  follows  that,  by  allow- 
ing 5,000  pounds  per  square  inch  of  section  at  c  d,  Fig.  233,  the  ten- 
sile stress  cannot  exceed  10,000  pounds  per  square  inch  at  a  b,  which  is  correct,  and 
agrees  with  practice. 

In  calculating  the  strength  of  the  key,  we  may  assume,  and  do  so  without  fear  of 
error,  that  the  tensile  stress  is  equal  to  the  shearing  stress.  The  key  is  subjected  to  a 
double  shear;  that  is,  shearing  must  take  place  at  two  places  before  fracture  can 
occur.  Hence,  the  area  of  one  cross-section  of  the  key  must  be  equal  to  one-half  the 
area  at  a  &  in  Fig.  233. 

TABLE   14. 


-7; 


Fig.  233 


Diameter  of  Piston. 

Diameter  of  Piston-rod. 

Large  Diameter  of  Tapered  End. 

9  inches. 

1^  inches. 

H  inches. 

10 

U 

1* 

11 

2 

If 

12 

3* 

H 

13 

2± 

o 

14 

2| 

2i 

15 

2* 

« 

16 

2* 

3* 

17 

2J 

21 

18 

3 

2J 

19 

3* 

« 

20 

3± 

3 

22 

3f 

3J 

LOCOMOTIVE  CONSTRUCTION. 


149 


The  dimensions  in  the  foregoing  table  agree  with  those  given  in  the  illustrations. 
The  principal  dimensions  of  pistons  are  given  in  Figs.  220  to  232. 

CROSSHEADS. 

106.  The  function  of  a  crosshead  is  to  form  a  connection  between  the  piston-rod 
and  the  connecting-rod,  making  the  piston-rod  move  in  its  true  course,  in  a  straight 
line,  while  the  connecting-rod  moves  through  various  oblique  positions.  Consequently 
we  may  say  the  crosshead  consists  essentially  of  a  socket,  to  which  the  piston  is 
keyed ;  a  journal,  on  which  one  end  of  the  connecting-rod  works ;  and  lastly,  sliding 
surfaces,  which  are  compelled  to  remain  in  contact  with  the  guides,  and  thus  guiding 


front  end. 


(I 


Fig.  236 


MISSION  SUITABLE  FOR  CAST  IRON. 

Cylinder  It'diumeter  *  gjf'ctrofce. 

the  end  nf  the  pisloii-n>d  in  its  true  course,  preventing  the  thrust  of  the  connecting- 
rod  from  bending  or  injuring  the  former. 

It  will  hardly  seem  necessary  to  give  names  to  such  familiar  mechanism  as  here 
represented ;  but  since  the  different  pieces  are  not  named  alike  by  all  mechanics,  the 
writer  deems  it  advisable  to  name  the  pieces,  so  as  to  avoid  misunderstanding  here- 
after. Similar  letters  in  the  different  views  represent  the  same 
piece. 

Figs.  234,  235,  and  236.  S  represents  the  guides,  frequently 
called  slides  (some  mechanics  call  these  the  slide-bars  or  guide- 
bars  ;  we  shall  simply  name  these  the  guides  or  slides) ;  Z?,  the 
guide-blocks ;  c,  the  crosshead ;  «?,  the  crosshead  wings,  or  simply 
the  wings  (in  some  books  these  are  called  slides  or  slide-blocks) ; 

<7,  the  gibs ;   P,  the  crosshead  pin ;  k,  the  crosshead  key ;  and  y, 

Fig.  2J5 
the  guide-yoke. 

In  locomotive  construction  three  styles  of  i-msslieads  are  used,  and  these  we  may 
classify  as  follows : 


150  MODERN   LOCOMOTIVE    CONSTRUCTION. 

First,  crossheads  which  require  four  guides;  second,  crossheads  which  require 
two  guides ;  third,  crossheads  which  require  one  guide. 

167.  Fig.  234  represents  a  side  view,  Fig.  235  an  end  view,  and  Fig.  236  a  plan  of 
a  crosshead  and  guides.  This  crosshead,  as  will  be  seen,  requires  four  guides.  This 
style  of  crosshead  is  generally  used  (not  always)  in  eight-wheeled  locomotives  such  as 
shown  in  Fig.  1,  and  ten-wheeled  locomotives  as  shown  in  Fig.  3.  When  a  crosshead 
is  to  be  designed  for  one  of  these  engines,  we  must  keep  in  mind  the  following  con- 
ditions. In  the  eight-wheeled  and  ten-wheeled  locomotives  the  rear  truck  wheels 
are  situated  directly  behind  the  cylinder  saddle  and  between  the  guides  and  the  frames, 
and  these  wheels  must  have  a  sufficient  space  for  the  lateral  play  when  the  engine  is 
running  over  a  curve.  Now  this  space  is  limited  by  the  position  of  the  cylinders. 
Thus,  for  instance,  the  centers  of  the  cylinders  are  always  placed  as  close  to  the  frames 
as  possible,  which,  of  course,  will  limit  the  space  between  the  guides  and  frames  and 
often  give  an  insufficient  space  for  the  lateral  play  of  the  rear  truck  wheels.  But, 
since  it  is  always  of  great  importance  to  keep  the  centers  of  cylinders  as  close  to  the 
frames  as  can  be  done,  we  should  not  attempt  to  spread  the  cylinders  so  as  to  obtain 
sufficient  space  for  the  wheels  until  allot  her  resources  fail.  Now  let  us  consider 
the  distance  between  the  center  of  the  cylinders  and  the  top  of  the  track,  and  see 
if  by  some  means  sufficient  room  can  be  obtained  for  the  wheels  to  pass  under- 
neath the  guides  and  thus  gain  for  them  a  sufficient  space  for  their  lateral  play. 
In  the  first  place,  the  centers  of  cylinders  are  usually  placed  from  1  to  2  inches 
above  the  centers  of  di'iving  wheels  when  the  engine  is  in  the  ordinary  good  running 
condition. 

The  center  of  the  crosshead  pin  P  must,  of  course,  remain  in  line  with  the  center 
of  the  piston-rod,  or — what  amounts  to  the  same  thing — the  center  line  of  the  cylinder ; 
hence  the  height  above  the  track  to  the  center  of  the  cylinder  or  the  crosshead  pin  P 
will  also  to  a  great  extent  limit  the  space  we  wish  to  obtain.  Now,  only  one  more 
resource  remains  by  which  we  may  obtain  the  desired  space,  and  that  is,  raising  the 
guides  above  the  center  of  the  crosshead  pin  P.  If  this  fails  to  give  us  sufficient  space, 
then  we  must  either  spread  the  centers  of  the  cylinders,  or  adopt  a  crosshead  such  as 
is  shown  in  Fig.  241,  or  do  both.  Usually  it  is'fouud  that  a  crosshead  with  the  guides 
raised  above  the  center  of  pin  P,  as  shown  in  Fig.  234,  will  answer  the  purpose  and 
still  allow  the  center  of  the  cylinders  to  remain  as  close  to  the  frames  as  they  can 
possibly  be  placed.  This  arrangement  will  enable  the  truck  wheels  to  pass  under- 
neath the  guides  as  far  as  will  be  necessary.  Here  the  flanges  have  not  entered  into 
our  consideration,  and  indeed  it  is  not  necessary  that  they  should,  as  there  will 
always  be,  in  this  class  of  engines,  sufficient  space  between  the  frames  and  guides  for 
the  lateral  play  of  the  flanges.  Sometimes  in  narrow-gauge  engines  the  side  of  the 
frames  must  be  cut  in  order  to  clear  the  truck  wheel.  It  must  also  be  remarked  that 
the  engine  and  truck  springs,  particularly  the  latter,  will  cause  the  cylinders  and  guides 
to  move  up  and  down,  and  since  the  wheels  have  not  this  movement,  a  sufficient 
clearance  between  the  under  side  of  the  guides  and  top  of  truck  wheels  must  be 
allowed,  to  prevent  the  guides  from  striking  the  top  of  truck  wheels  when  the  springs 
impart  a  vertical  movement  to  the  guides. 

This   style  of  crosshead,  Fig.  234,  is  usually  made  of  cast-iron,  sometimes  of 


MODERN  LOCOMOTITE   COXSTBl'CTION. 


151 


malleable  iron.  The  crosshead  pin  P  and  the  crosshead  are  cast  in  one  piece.  The 
pin  /'  is  generally  finished  off  by  hand,  sometimes  with  special  machinery. 

In  order  to  take  up  quickly  and  readily  any  wear  between  the  guides  and  the 
wings  of  the  crosshead,  the  brass  gibs  g  are  introduced.  The  lips  at  the  ends  of  these 
gibs  prevent  them  from  moving  endways,  and  the  lugs  h  cast  to  the  gibs  and  fitting  in 
the  slots  cut  in  the  wings  of  the  crosshead  will  prevent  the  gibs  from  slipping  out 
sideways.  When  the  wear  is  to  be  taken  up,  thin  copper  liners  are  inserted  between 
the  gibs  and  the  wing. 

168.  Some  master-mechanics  prefer  crossheads  with  brass  gibs,  because,  the  brass 
being  softer  than  the  iron,  they  believe  that  there  is  not  so  great  a  liability  to  cut  the 
guides  as  when  cast-iron  bears  directly  against  them,  which  will  be  the  case  when 
crossheads  without  the  gibs  are  used.  Cutting  the  guides  is  always  a  serious  matter, 


Q. 

Q.     D- 

S 

r 

P=  —  ^ 

;=> 

Q. 

0- 

D« 

Fip.  252 

mg.  254, 


Fig.  253 


and  if  cutting  does  occur,  it  is  always  preferable  that  the  gibs  should  be  cut  or  ruined, 
rather  than  the  guides,  as  the  former  are  cheaper  to  replace,  and  can  be  replaced  in  less 
time  (another  important  matter  in  railroading)  than  ruined  guides.  Yet  experience 
also  teaches  that  when  guides  are  properly  case  hardened,  and  crossheads  without  gibs 
are  used,  so  that  cast-iron  bears  directly  against  the  guides,  cutting  of  the  guides  is 
prevented  by  running  a  new  engine  at  first  slowly  and  carefully,  allowing  the  cast- 
iron  wings  to  wear  down  to  smooth,  hard  surfaces ;  after  that  there  is  little  danger  of 
cutting  the  guides,  providing  they  are  kept  oiled.  Consequently  we  meet  with  many 
locomotives  in  which  the  crossheads  have  no  brass  gibs  and  give  perfect  satisfaction. 

Crossheads  without  gibs  have  the  wings  babbitted,  as  shown  at  a  a  a  in  Fig.  252. 
Three  or  more  rectangular  recesses,  about  £  or  f  of  an  inch  deep,  f  of  an  inch  wide, 
and  as  long  as  the  width  of  the  wing  will  allow,  are  cast  in  the  wing,  and  then  filled 
with  Babbitt  metal.  Sometimes  as  many  recesses,  \  or  f  of  an  inch  deep,  Ij  inches 
in  diameter,  as  there  is  room  for,  arranged  in  a  manner  as  shown  at  b  b  b  in  Fig.  253, 
are  bored  in  the  wing,  and  then  filled  with  Babbitt  metal.  A  few  years  ago  these 
recesses  were  sometimes  filled  with  glass  disks.  When  these  were  used,  the  bottom 
diameter  of  the  recess  was  larger  than  the  upper  diameter,  and  also  the  lower  diameter 
of  the  glass  disk  larger  than  the  upper  one.  This  disk  was  then  placed  in  the  recess,  and 
Babbitt  metal  poured  around  it,  as  shown  in  Fig.  254,  and  thus  held  firmly  in  the  recess. 

169.  Objections  are  sometimes  raised  against  crossheads  with  the  pins  P  cast  in 
one  piece,  as  it  is  difficult  to  true  up  these  pins  when  necessary,  and  therefore  cross- 
heads  similar  to  that  as  represented  in  Figs.  237,  238,  239,  and  240  are  sometimes 
preferred. 


152 


MOI>KI;\  LOCOMOTIVE  <:o.\xTi;r<"n<>\. 


This  crosshead  consists  of  several  pieces.  Fig.  240,  His  a  wrought-iron  hub  with 
fork  end.  This  hub  is  keyed  to  the  piston-rod.  The  wings  w3  w3,  with  flanges  13  13, 
are  made  of  cast-iron.  The  bushing  F  is  made  of  wrought-iron  case  hardened,  the 
grain  of  the  iron  running  around  the  bushing,  and  not  in  the  direction  of  its  length ; 
the  fork  end  of  the  hub  H  is  bored  out  large  enough  to  receive  the  bushing.  The  bolt 
«3  is  made  of  wrought-iron  not  case  hardened ;  the  outer  plate  «3  is  made  of  brass  and 


LOCOMOTIVE  CROSSHEAD 

't  n 

Cylinders  17  diatn.  84  stroke. 


6  «iJS     * .,  * 


Cylinders  If  diam.  £  34  stroke. 


the  inner  plate  b3  of  cast-iron.  The  purpose  of  the  holes  r3  r3  in  the  wing  is  to  reduce 
the  weight  of  the  latter.  Thin  brass  plates  d  d  (Fig.  239)  are  riveted  to  the  bearing 
surfaces  of  the  wing,  as  shown.  In  putting  this  crosshead  together,  the  bolt  o3  (Fig. 
240)  is  passed  through  the  cast-iron  plate  13,  and  through  the  wings,  with  the  hub  H 
and  bushing  F  between  them.  In  screwing  these  together,  the  wings  w3  w3  bear  hard 
against  the  ends  of  the  bushing  F,  and  thus  prevent  the  closing  up  of  the  fork  on 
the  hub  H.  The  brass  plate  a3  is  then  fastened  to  the  outer  wing  by  two  f  screw 
bolts,  covering  up  the  nut  of  the  bolt  o3  and  giving  a  nice  and  clean  outer  appearance 
to  the  crosshead.  The  f  screw  bolt  n3  prevents  the  cast-iron  plate  from  turning  on 
the  bolt  o3.  The  bushing  F  forms  the  crosshead  pin,  and  when  it  becomes  worn,  can 
easily  be  removed  and  replaced.  This  design  makes  a  very  good  crosshead,  but  on 
account  of  its  expense  is  not  often  used. 

When  crossheads  of  this  kind  or  those  without  gibs  are  used,  thin  copper 
strips  are  inserted  between  the  guide-blocks  B  B  and  the  guides  s  s  at  the  time  the 
engine  is  being  built.  Then  when  it  becomes  necessary  to  take  up  the  wear  between 


MODEKX  LOCOMOTIVE   CONSTRUCTION.  153 

the  guides  and  the  crosshead,  these  copper  slips  are  one  by  one  removed,  thereby 
bringing  the  guides  together. 

170.  The  pressure  of  locomotive  crossheads  against  the  guides  caused  by  the  thrust 
of  the  connecting-rod  should  not  exceed  50  pounds  per  square  inch ;  and  how  to  find 
this  thrust  we  will  presently  explain.    From  this  remark  we  conclude  that  the  sliding 
surfaces  of  the  wings  of  crossheads.  shown  in  Figs.  234  and  237  must  contain  a  certain 
number  of  square  inches,  and  consequently  if  the  width  of  the  guides  is  increased  the 
length  of  the  wings  is  made  less,  and  when  the  length  of  the  wings  is  shortened 
the  guides  are  also  shortened,  which  is  always  desirable.    When  four  guides  are 
employed,  as  shown  in  Figs.  234  and  237,  we  naturally  obtain  a  wide  sliding  surface, 
and  consequently  the  crossheads  and  guides  shown  in  these  figures  are  comparatively 
short ;  the  guides,  being  placed  well  up  above  the  rails,  are  kept  comparatively  free 
from  dust ;  the  crosshead  is  light ;  in  fact,  the  whole  arrangement  is  well  adapted  for 
eight-wheeled  passenger  engines,  or  engines  having  a  four-wheeled  truck  in  front. 

The  vertical  distance  between  the  guides,  as  shown  in  Figs.  234  and  237,  should  be 
only  sufficient  to  admit  a  wing  of  a  minimum  depth.  In  these  figures  it  will  also 
be  noticed  that  the  crosshead  pin  P  is,  horizontally,  somewhat  out  of  the  center  of  the 
wings  or  sliding  surfaces  of  the  crosshead.  The  reason  for  this  is  that  designers  will 
always  endeavor  to  make  the  crosshead,  and  consequently  the  guides,  as  short  as 
possible.  Now,  because  the  required  strength  of  the  crosshead  will  fix  the  distance 
between  the  center  of  the  pin  P  and  the  front  end  of  hub,  this  distance  is  limited,  but 
by  moving  the  wings  a  little  ahead  of  the  pin  P,  which  can  often  be  done,  the  distance 
from  the  front  of  the  hub  to  the  rear  end  of  the  wing  is  somewhat  decreased,  and 
therefore  the  guides  can  also  be  made  a  little  shorter.  But  in  the  writer's  opinion 
such  practice  should  be  avoided  as  much  as  possible,  and  the  pin  P  should  be  kept 
central  with  the  wings  or  sliding  surfaces  of  ah1  crossheads ;  by  so  doing  the  latter 
will  wear  more  evenly. 

171.  The  type  of  crosshead  of  which  a  side  elevation  is  shown  in  Fig.  241,  an  end 
elevation  in  Fig.  242,  and  a  plan  in  Fig.  243,  is  occasionally  used  for  eight-wheeled 
passenger  engines,  but  the  writer  believes  that  this  crosshead  is  better  adapted  for  a 
mogul  engine,  such  as  is  shown  in  Fig.  2,  and  a  consolidation  engine,  such  as  is  shown 
in  Fig.  4,  and  for  these  engines  it  is  very  often  used.     This  design  of  crosshead  is 
suitable  for  cast-iron,  of  which  these  crossheads  are  made,  and  the  dimensions  here 
given  are  suitable  for  a  cylinder  20  inches  in  diameter.     The  crosshead  pin  P  is  either 
made  of  steel  or  wrought-iron  case  hardened.     The  head  of  the  pin  is  always  placed 
towards  the  driving  wheels ;  by  so  doing — as  will  be  obvious — the  cylinders  can  be 
brought  closer  together  and  still  leave  room  enough  for  the  crosshead  to  pass  the 
crank-pin  of  the  front  driving  wheels.     The  side  of  the  crosshead  in  which  the  head 
of  the  pin  P  is  inserted  we  shall  lien-after  call  the  inner  side  of  the  crosshead,  and  the 
opposite  one  the  outer  side.     In  the  majority  of  crossheads  of  this  type  brass  gibs 

•<1i'.l-i  (Fig-  242)  are  used.  In  order  to  take  these  out  quickly  when  necessary,  or  to 
put  in  liners  when  the  wear  demands  it,  the  outer  flanges  f2f2  are  bolted  to  the  cross- 
head  ;  the  inner  flanges  are  cast  to  it.  These  flanges  and  the  lips  cast  on  the  ends  of 
the  gibs  will  prevent  the  hitter  from  slipping  out  of  place.  In  designing  a  crosshead  one, 
great  object  aimed  at  is  to  make  it  as  light  as  possible,  and  still  leave  it  sti-ong  enough 


154 


MODERN   LOCOMOTIVE    CONSTRUCTION. 


-Ifc 


DESIGN  SUITABLE  FOR  CAST  IRON 


DESIGN  SUITABLE  FOR  WROUGHT  IRON  OR  CAST  STEEL. 

Cylinder  «O  inches  diameter.  34  Inches  stroke. 


MODERN  LOCOMOTIVE  CONSTRUCTION.  155 

to  meet  any  emergency ;  consequently  the  distance  between  the  guides  must  be  as  short 
as  possible.  This  distance  is  determined  by  the  oblique  positions  of  the  connecting- 
rod.  The  method  employed  for  finding  the  distance  between  the  guides  will  be 
explained  hereafter.  Some  master-mechanics  raise  an  objection  to  this  type  of  cross- 
head,  because  when  it  is  used  in  engines  which  have  driving  wheels  of  comparatively 
small  diameter,  as  is  the  case  in  freight  engines,  the  guides  are  brought  too  close  to 
the  rails  and  consequently  exposed  to  the  dust,  which  will  wear  the  guides  and  cross- 
Ix-tid  too  fast.  To  avoid  this  difficulty,  crossheads  are  used  of  which  Fig.  244  is  a  side 
elevation,  Fig.  246  an  end  elevation,  Fig.  247  a  plan,  and  Fig.  245  a  view  of  the  cross- 
head  pin  P.  This  design  of  crosshead  is  adapted  for  wrought-iron  or  cast  steel,  and  is 
made  of  either  one  or  the  other  material.  The  dimensions  here  given  are  suitable  for 
cylinders  20  inches  in  diameter  with  a  maximum  steam  pressure  of  120  pounds  per 
square  inch.  Some  types  of  wrought-iron  crossheads  are  very  expensive  to  make, 
and  if  wrought-iron  is  insisted  upon,  then  this  design  recommends  itself,  as  the 
expense  connected  with  making  a  crosshead  of  this  type  is  comparatively  small. 
Referring  to  Figs.  244  and  246  it  will  be  seen  that  this  crosshead  consists  simply  of  a 
hub  H  H2,  to  which  the  piston-rod  is  keyed.  To  this  hub  are  forged  two  deep  flanges 
F  F2.  These  flanges  extend  upwards  and  terminate  a  little  below  the  upper  guides 
S  S2.  Between  these  flanges  a  cast-iron  block  h  h2  is  bolted  by  a  number  of  bolts  £  of 
an  inch  in  diameter ;  these  bolts  extend  through  the  flanges  and  the  cast-iron  block. 
To  the  upper  and  lower  faces  of  the  block  Ji  h2  the  brass  gibs  g  g2  g2  are  fitted.  The 
gibs  are  held  in  position  sideways  by  the  flanges  F2  F2,  and  endways  by  the  lips  cast 
to  the  gibs ;  for  additional  safety  two  pins  r  r  one  inch  in  diameter  are  driven  through 
the  gibs  g  g  and  the  block  h.  The  flanges  of  the  upper  gib  g.2  slide  along  the  sides  of 
the  upper  guide,  and  the  flanges  F2  F2  slide  along  the  sides  of  the  lower  guides, 
thus  forming  good  deep  sliding  surfaces  which  will  guide  the  end  of  the  piston-rod  in 
a  straight  line  laterally,  although  not  with  such  steadiness  as  the  crosshead  shown  in 
Fig.  241  will  do ;  and  in  this  respect  the  crosshead  represented  in  Fig.  244  is  somewhat 
inferior  to  the  one  shown  in  Fig.  241.  The  gibs  in  the  crosshead  shown  in  Fig.  244 
are  not  so  readily  removed  as  in  those  shown  previously,  because  in  the  case 
before  us,  in  order  to  take  out  the  gibs  the  upper  guide  must  be  taken  off,  besides 
taking  out  all  the  bolts  which  hold  the  block  7i,  and  this  is  not  always  an  easy  matter. 
The  distance  from  the  center  of  the  crosshead  pin  P  to  the  lower  surface  of  the 
bottom  guide  must  be  as  short  as  possible ;  it  should  be  only  sufficient  to  allow  the 
connecting-rod  when  in  an  oblique  position  to  clear  the  edge  o  of  the  lower  guide. 

172.  Fig.  248  is  a  side  elevation,  Fig.  250  an  end  elevation,  and  Fig.  251  a  plan 
of  another  crosshead  and  guide ;  and  Fig.  249  shows  the  crosshead  pin  P. 

This  crosshead  is  made  of  cast-iron,  and,  as  will  be  seen,  requires  only  one  guide. 
This  design  of  crosshead  should  only  be  used  for  small  locomotives ;  the  dimensions 
here  given  are  suitable  for  a  cylinder  15  inches  in  diameter,  and  indeed  this  crosshead 
is  seldom  used  on  engines  having  cylinders  larger  than  15  inches  in  diameter.  This 
crosshead,  with  the  exception  of  the  plate  F  F2  which  is  bolted  on,  is  cast  in 
one  piece;  this  arrangement  allows  the  brass  gibs  to  be  readily  placed  in  position, 
and  quickly  removed  when  necessary.  The  distance  between  the  guide  and  the 
pin  P  must  be  such  as  will  allow  the  end  0  of  the  guide  to  clear  the  connecting-rod 


156 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


when  in  an  oblique  position.  The  writer  believes  this  crosshead  to  be  inferior  to 
all  the  others  shown,  because  in  his  opinion  it  will  not  guide  the  end  of  the  piston-rod 
laterally  as  steadily  as  the  others.  Yet  it  is  a  very  cheap  crosshead,  and  seems  to 
work  well  in  small  engines. 


TO   FIND   THE   THRUST   OF   THE   CONNECTING-ROD. 

173.  When  the  connecting-rod  stands  in  an  oblique  position,  for  instance  such  as 
will  occur  when  the  crank  is  at  half  stroke,  the  connecting-rod  will  force  the  crosshead 


Fig.  250 


Fig.  249 


Inner  side, 


Fig.  251 


DESIGN  SUITABLE  FOR  CAST  IKON. 

Cylinder  IS  incites  diameter.  SO  inches  stroke. 


against  the  guides ;    and  this  force  which  presses  the  crosshead  against  the  guides  is 
called  the  thrust  of  the  connecting-rod.     The  amount  of  this  thrust,  or  the  magnitude 
of  this  force,  can  be  found  by  two  methods :  The  graphic  method,  and  by  calculation. 
We  will  first  explain  how  this  thrust  can  be  found  by  the  graphic  method. 


MODERN  LOCOMOTIVE   CONSTRUCTION.  157 

Fig.  252A.  Let  D  be  the  center  of  the  driving  axle,  C  the  center  of  the  crank-pin, 
P  the  center  of  the  crosshead  pin,  S  S  the  guides,  D  I  the  center  line  of  motion,  and 
the  circumference  R  C  the  path  of  the  center  of  crank-pin.  We  shall  assume  that 
steam  follows  the  full  stroke  of  piston.  It  is  required  to  find  the  thrust  of  the  con- 
necting-rod, or  the  pressure  of  the  crosshead  against  the  guides. 

The  total  steam  pressure  on  the  piston  is  transmitted  through  the  piston-rod  to 
the  crosshead  pin  P,  and  from  thence  it  is  transmitted  through  the  connecting-rod  to 
the  crank-pin  C.  The  directions  in  which  this  steam  pressure  acts  is  in  the  direction 
of  the  center  line  of  the  piston-rod,  and  in  the  direction  of  the  center  line  of  the  con- 
necting-rod. 

At  the  beginning  of  a  stroke,  the  center  line  of  the  piston-rod  and  the  center  line 
of  the  connecting-rod  will  lie  in  one  and  the  same  straight  line,  and  therefore  we 
assume  that  in  this  position  there  will  be  no  thrust,  or,  in  other  words,  the  steam 
pressure  will  not  cause  any  pressure  between  the  crosshead  and  the  guides. 

When  the  crank  stands  perpendicular  to  the  center  line  of  motion  I)  b,  we  may 
assume  that  for  the  purpose  of  finding  the  thrust  the  angle  formed  by  the  line  D  b 
and  the  center  line  C  P  of  the  connecting-rod  will  be  the  greatest,  and  therefore  we 
conclude  that  in  this  position  the  thrust  of  the  connecting-rod  will  also  be  the  greatest. 
It  is  now  our  object  to  find  the  intensity  of  the  thrust,  or  the  magnitude  of  this  force, 
when  the  crank  D  C  stands  at  right  angles  to  the  center  line  of  motion  D  b. 

Draw  the  center  line  of  motion  D  b ;  on  this  line  take  any  point,  as  D,  to  repre- 
sent the  center  of  the  driving  axle ;  through  the  point  D  draw  a  line  D  C  perpendicu- 
lar to  the  line  D  b,  and  make  D  C  equal  to  the  length  of  the  crank.  From  C  as  a 
center,  and  with  a  radius  equal  to  the  length  of  the  connecting-rod,  draw  a  short  arc 
cutting  the  line  D  b  at  the  point  P ;  this  point  will  be  the  center  of  the  crosshead  pin, 
and  the  line  C  P  will  represent  the  center  line  of  the  connecting-rod.  Prolong  the 
line  C  P  to  e,  making  P  e  equal  to  C  P.  Through  the  point  P  draw  a  line  P  a 
perpendicular  to  the  line  D  b ;  through  the  point  e  draw  a  line  e  b  parallel  to  P  a, 
and  again  through  the  point  e  draw  a  line  e  a  parallel  to  P  ft ;  then  P  a  e  b  will  be  a 
parallelogram,  which  is  called  the  parallelogram  of  forces. 

Now  in  mechanics,  which  is  that  branch  of  science  which  treats  of  the  effects  of 
forces  upon  matter,  it  is  shown  that  forces  can  be  completely  represented  by  straight 
lines,  or,  in  other  words,  the  magnitude  of  a  force  and  the  direction  in  which  it  acts 
can  be  represented  by  a  straight  line.  It  is  also  further  shown  that  in  a  parallelogram 
of  forces,  such  as  we  have  just  completed,  the  magnitudes  of  the  forces  are  pro- 
poi-tional  to  the  lengths  of  the  sides  of  the  parallelogram  of  forces ;  that  is  to  say,  if 
the  length  of  the  side  P  a  in  our  parallelogram  is  equal  to  £  the  length  of  the 
side  P  b,  then  the  force  represented  by  the  side  P  a  will  be  one  half,  of  the  force 
represented  by  the  side  P  b ;  or  again,  if  the  length  of  the  line  P  a  is  equal  to  £  of  the 
length  of  P  fc,  then  the  force  represented  by  the  line  P  a  will  be  one  quarter  of  the 
force  represented  by  the  line  or  side  P  b. 

Here,  then,  it  may  be  said  that  we  have  a  point  P  which  is  held  in  equilibrium  by 
three  forces.  The  magnitude  of  these  forces  and  the  direction  in  which  they  act  are 
completely  represented  by  the  length  and  the  direction  of  the  three  straight  lines  C  P, 
P  b,  and  P  a.  The  line  P  b  represents  the  total  steam  pressure  on  the  piston,  which  is 


158  MODERN  LOCOMOTIVE   CONSTRUCTION. 

acting  in  the  direction  of  the  center  line  of  the  piston-rod ;  the  line  C  P  represents  a 
force  acting  in  the  direction  of  the  center  line  of  the  connecting-rod ;  and  the  line  P  a 
represents  a  force  acting  in  a  direction  perpendicular  to  the  guides,  and  is  the  thrust 
of  the  connecting-rod,  or,  in  other  words,  the  line  P  a  represents  the  magnitude  of  the 
force  which  presses  the  crosshead  against  the  guides,  and  which,  according  to  our 
problem,  was  to  be  found. 

174.  In  order  to  find  the  force  or  thrust  P  a  it  is  not  necessary  to  draw  the  cross- 
head,  guides,  piston-rod,  and  as  many  lines  as  we  have  done  in  this  figure  (this  was 
simply  done  to  make  the  principles  plain),  but  we  can,  without  adding  any  new 
principles  or  changing  the  foregoing  ones,  obtain  the  same  forces  by  constructing  the 
triangle,  Fig.  253 A,  which  may  be  drawn  full  size,  half  size,  or  to  any  convenient  scale 
thus: 

Draw  any  straight  line  P  b  (Fig.  253A) ;  on  this  line  take  any  point,  as  6,  and 
through  this  point  draw  a  line  b  e  perpendicular  to  the  line  P  I ;  on  the  line  b  e  lay 

off  a  point  e ;  the  distance  between 
the  points  b  and  e  must  be  equal 
to  the  length  of  the  crank.  From 
the  point  e  as  &  center,  and  with  a 

f  Fig.ZBZA  *  r  7 

radius  equal  to  the  length  of  the 

connecting-rod,  describe  a  short  arc  cutting  the  line  P  b  in  point  P ;  join  the  points  P 
and  e  by  a  straight  line,  and  complete  the  triangle  P  b  e.  Now,  if  the  dimensions 
of  engine  and  steam  pressure  in  Fig.  253A  remain  the  same  as  those  in  Fig.  252A, 
the  triangle  shown  in  Fig.  253A  will  be  equal  to  any  of  the  triangles  as  P  b  e,  P  a  e,  and 
P  D  C  in  Fig.  252A,  and  consequently  the  sides  of  the  triangle  in  Fig.  253A  will  repre- 
sent the  same  forces  as  the  sides  P  a,  a  e,  and  the  diagonal  P  e  of  the  parallelogram 
P  a  b  e  in  Fig.  252A. 

PRACTICAL  APPLICATIONS   OF  THE  FOREGOING  PRINCIPLES. 

175.  EXAMPLE  45. — Diameter  of  the  piston  is  16  inches,  the  stroke  is  24  inches, 
length  of  connecting-rod  84  inches,  and  the  steam  pressure  is  120  pounds  per  square 
inch ;  steam  follows  full  stroke :  find  the  thrust  of  the  connecting-rod  or  the  pressure 
of  the  crosshead  against  the  guides. 

Draw  the  line  P  b  (Fig.  253A) ;  at  any  point  b  on  this  line  erect  the  perpendicular 
b  e ;  make  the  length  of  e  b  equal  to  12  inches  (which  is  the  length  of  the  crank  or  half 
the  given  stroke).  From  e  as  a  center,  and  with  a  radius  equal  to  84  inches  (the 
given  length  of  connecting-rod),  describe  a  short  arc  cutting  the  line  P  b  at  the  point 
P ;  join  P  and  e  by  a  straight  line,  and  the  triangle  will  be  completed. 

The  total  steam  pressure  on  the  piston  is  found  by  multiplying  its  area  in  square 
inches  by  the  steam  pressure  per  square  inch,  and  this  total  pressure  will  be  equal  to 
24,120  pounds  (the  fraction  has  been  omitted) ;  hence  the  length  of  the  line  P  b  will 
represent  24,120  pounds. 

Now  assume  that  we  have  a  narrow  strip  of  paper  whose  length  is  exactly  equal 
to  the  length  of  the  line  P  b,  and  that  this  paper  is  divided  lengthwise  into  24,120 
equal  parts,  then  each  division  will  represent  one  pound ;  laying  this  piece  of  paper 
(or  scale,  as  we  may  call  it)  on  the  line  e  b,  we  find  that  this  line  will  contain  3,481 


MODERN  LOCOMOTirE   CONSTRUCTION. 

(iiciirly)  <>t'  the  number  of  divisions  on  the  paper;  hence  we  conclude  that  the  thrust, 
or  the  pressure  of  the  crosshead  against  the  guides,  is  3,481  pounds. 

But  to  divide  the  line  P  b  into  24,120  equal  parts  would  require  too  much  time 
and  labor,  hence  the  following  method  is  used : 

Let  us  adopt  4  of  an  inch  to  represent  1,000  pounds;  then  since  24,120  pounds  is 
represented  by  line  P  ft,  we  lay  off  from  the  point  P  on  the  line  P  b  twenty-four  J 
inches,  which  will  then  represent  24,000  pounds,  because  each  £  inch  represents  1,000 
pounds.  In  order  to  represent  the  remaining  120  pounds,  which  are  very  nearly  equal 
to  £  of  1,000  pounds,  we  must  add  &  of  £  inch  (which  is  equal  to  -^  of  an  inch)  to  the 
twenty-four  %  inches ;  or,  in  short,  from  the  point  P  on  the  line  P  b  lay  off  a  point  x ; 
the  distance  between  the  points  P  and  x  must  be  equal  to  12-^  inches.  Through  the 
point  x  draw  a  line  x  y  perpendicular  to  the  line  P  b  and  cutting  the  line  P  e  in  the 
point  y ;  then  the  line  x  y  will  represent  the  thrust  of  the  connecting-rod,  and  since 
the  length  of  this  line  is  very  nearly  equal  to  If  inches,  which  we  obtain  by  measure- 
ment, and  since  each  £  inch  represents  1,000  pounds,  we  know  that  the  pressure  of  the 
crosshead  against  the  guide  is  not  quite  but  very  nearly  equal  to  3,500  pounds ;  and 
this  answer  is  in  most  cases  near  enough  for  practical  purposes. 

But  should  it  be  necessary  to  find  the  amount  of  this  thrust  accurately,  then  the 
simplest  way  to  determine  it  is  by  calculation ;  and  if  the  foregoing  graphic  method 
is  understood,  then  there  should  not  be  any  difficulty  in  understanding  the  following 
calculations,  which  are  based  on  the  principles  already  introduced  in  connection  with 
the  graphic  method,  thus : 

176.  We  have  already  seen  that  the  magnitudes  of  the  three  forces  which  hold  the 
point  P  in  Fig.  252A  in  equilibrium  are  represented  by  the  three  sides  of  the  triangle 
shown  in  Fig.  253A.  Now  we  know  the  length  of  the  line  P  e,  which,  according  to  our 
example,  is  84  inches ;  we  also  know  the  length  of  the  line  b  e,  which  is  12  inches,  but 
the  length  of  the  line  P  b  we  do  not  know,  yet  we  do  know  that  the  length  of  this 
latter  line  must  represent  24,120  pounds.  If  we  now  find  by  calculation  (instead  of 
construction  as  before)  the  length  of  the  line  P  ft,  then,  since  we  know  the  lengths  of 
the  other  lines  or  sides  of  the  triangle  we  shall  have  no  difficulty  in  computing  the 
number  of  pounds  that  each  of  the  sides  of  this  triangle  represents,  because  the 
magnitudes  of  the  forces  are  proportional  to  the  length  of  the  lines. 

The  triangle  P  e  b  is  by  construction  a  right-angled  triangle,  and  consequently  to 
find  the  length  of  the  side  P  b  we  subtract  the  square  of  the  side  b  e  from  the  square 
of  the  side  P  e ;  the  square  root  of  the  remainder  will  be  the  length  of  the  side  P  b. 

The  length  of  P  e  is  84  inches,  the  length  of  b  e  is  12  inches.  The  square  of  84  is 
equal  to  84  x  84  =  7,056,  and  the  square  of  12  is  equal  to  12  x  12  =  144,  and  7,056  - 
144  =  6,912.  The  square  root  of  6,912  is  equal  to  83.13+  inches,  which  is  the  length 
of  the  side  P  b.  But  the  side  P  b  and  consequently  the  83.13  inches  represent  24,120 
pounds ;  therefore  the  number  of  pounds  of  the  force  represented  by  the  side  b  e,  which 
bears  the  same  proportion  to  24,120  pounds  as  the  length  of  b  e  bears  to  P  b,  is  found 
by  the  simple  rule  of  proportion,  thus : 

83.13  inches  :  12  indies  ::   24,120  :  the  answer; 
hence, 

24,120  x  12       .,  1U1 

go  -.o      —  ^MHI  pounds, 


160 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


Therefore  the  magnitude  of  the  force  represented  by  the  side  b  c  is  equal  to  3,481 
pounds,  which  is  the  pressure  of  the  crosshead  against  the  guides.  Now  putting  the 
whole  foregoing  calculations  in  the  shape  of  a  formula,  we  have  the  following : 


FORMULA  AND   KULE  FOR  FINDING   THE   THRUST   OF    THE   CONNECTING-ROD   BY   CALCULATION. 

Length  of  crank  in  inches 
Total  pressure  of  steam  on  piston  x  ^  (length  of  COnneeting-rod  in  inches)2  -  (tentfh~of"^Snk"in  inches)1 

=  thrust  of  connecting-rod. 

• 

Or,  in  ordinary  language,  we  have  the  following : 

RULE  19. — Multiply  the  total  steam  pressure  in  pounds  on  the  piston  by  the  length 
of  the  crank  in  inches,  and  call  this  product  a ;  then  from  the  square  of  the  length  of 
the  connecting-rod  in  inches  subtract  the  square  of  the  length  of  the  crank  in  inches, 
and  find  the  square  root  of  the  remainder ;  divide  the  product  a  by  the  last  answer, 
the  quotient  will  be  the  thrust  of  the  connecting-rod  in  pounds. 

177.  The  foregoing  rule  will  give  the  thrust  of  the  connecting-rod  correctly  in 
stationary  and  marine  engines,  or  in  all  engines  in  which  the  center  I)  of  the  crank- 
shaft (Fig.  252A)  cannot  move  out  of  the  center  line  of  motion  D  b ;  but  such  conditions 
do  not  exist  in  locomotives,  hence  the  foregoing  rule  must  be  somewhat  modified,  so 
that  it  will  apply  to  them. 

Fig.  254A.  Let  D  b  represent  the  center  line  of  motion ;  I),  the  center  of  driving 


*'»(/.  254  A 


axle ;  C,  the  center  of  crank-pin ;  the  circumference  E  C,  the  path  of  the  center  of  crank- 
pin  ;  and  P,  the  center  of  the  crosshead-pin. 

In  locomotives  the  driving-axle  box  can  move  up  and  down  in  the  pedestals,  and 
consequently  the  center  7),  and  with  it  the  path  E  C  of  the  center  of  crank-pin,  will 
move  out  of  the  center  line  of  motion  D  b. 

In  order  to  obtain  the  greatest  thrust  of  a  connecting-rod  in  a  locomotive,  we  find 
the  extreme  lower  and  upper  position  C*2  and  C3  of  the  crank-pin  when  the  center  line 
D  <72  of  the  crank  stands  perpendicular  to  the  center  line  of  motion  D  b.  Now  let  C2 
represent  the  center  of  crank-pin  in  its  extreme  lower  position;  then  in  order  to 
construct  the  parallelogram  of  forces  we  proceed  in  a  manner  similar  to  that  adopted 
before,  namely: 

From  the  point  6'2  as  a  center,  and  with  the  length  of  the  connecting-rod  as  a 
radius,  describe  a  short  arc  cutting  the  center  line  of  motion  D  b  at  the  point  P,  which 


MODERN  LOCOMOTIVE   CONSTRUCTION. 

will  be  the  position  of  the  center  of  crosshead-pin  when  the  crank  stands  at  right 
angles  to  the  center  line  of  motion  and  the  crank-pin  in  its  extreme  lowest  position. 
From  C.,  and  through  P  draw  a  straight  line,  and  produce  it  towards  e ;  make  the  line 
P  e  equal  in  length  to  that  of  C2  P,  or,  in  other  words,  to  the  length  of  the  connecting- 
rod.  Through  the  point  e  draw  a  straight  line  e  a  parallel  to  P  6,  and  through  the 
points  /'  and  e  draw  the  lines  P  a  and  b  e  perpendicular  to  P  b,  thus  completing  the 
parallelogram  of  forces  P  a  b  e. 

Now,  if  in  Fig.  254A  the  length  of  the  connecting-rod,  crank,  and  the  total  steam 
pressure  is  the  same  as  those  in  Fig.  232A,  the  line  b  e  in  Fig.  254A  must  be  considerably 
longer  than  b  e  in  Fig.  252A,  and  therefore  the  thrust  represented  by  the  line  b  e  in  the 
former  figure  must  be  greater  than  the  thrust  in  the  latter  figure.  Consequently  the 
formula  for  determining  the  thrust  of  a  locomotive  connecting-rod  will  be  as  follows : 

BULE  20,   IN   SYMBOLS. — TO    FIND   THE   THRUST   OF   A  LOCOMOTIVE  CONNECTING-ROD. 

Let  P  represent  the  total  pressure  of  the  steam  in  pounds  on  the  piston ;  L,  the 
length  in  inches  of  a  line  drawn  perpendicular  to  the  center  line  of  motion  D  b  and 
measured  from  the  line  D  b  to  the  extreme  lowest  position  of  the  center  C  of  the 
crank-pin ;  K,  the  length  of  the  connecting-rod  in  inches ;  T,  the  thrust  of  the  connect- 
ing-rod, or  the  pressure  of  the  crosshead  against  the  guides  in  pounds.  Then 

P  x  — =     L      =  =  T. 
V  K*  -  L* 

If  the  line  D  C3  should  be  greater  than  D  C2,  then  the  thrust  will  be  the  greatest 
when  the  crank-pin  is  in  its  extreme  upper  position,  and  in  the  calculation  the  length 
of  D  Cy  should  be  taken  in  place  of  D  C2. 


APPROXIMATE  RULE  FOE  FINDING  THE  THRUST  OF  A  CONNECTING-ROD. 

178.  When  the  length  of  the  connecting-rod  is  very  great  in  comparison  with  that 
of  the  crank,  then  the  difference  between  the  lines  P  e  and  P  b  is  very  slight,  and 
for  many  practical  purposes  this  difference  may  be  neglected.  In  such  cases  we 
assume  that  the  length  of  the  line  P  e  will  represent  the  total  pressure  on  the  piston, 
and  consequently  the  rule  for  finding  the  thrust  becomes  very  simple. 

APPROXIMATE  RULE  21. — Divide  the  length  of  the  crank  in  inches  by  the  length  of 
the  connecting-rod,  and  multiply  the  quotient  by  the  total  steam  pressure  on  the 
piston ;  the  product  will  be  the  thrust  of  the  connecting-rod  against  the  guides,  in 
pounds. 

Applying  this  rule  to  Example  45,  we  have  the  following  result : 

It  will  be  remembered  that  the  total  steain  pressure  on  the  piston  was  found  to 
be  equal  to  24,120  pounds,  the  length  of  the  connecting-rod  is  84  inches,  and  the 
length  of  the  crank,  12  inches.  Hence 

if  x  24,120  =  3,445.71+  pounds  =  thrust  of  the  connecting-rod. 
The  thrust  of  the  connecting-rod  found  by  Rule  1(J  was  3,481  pounds.     The  dif- 


162  MODERN  LOCOMOTIVE   CONSTRUCTION. 

ference  between  the  two  answers  is  so  slight  that  it  may  be  neglected,  and  in  cases  of 
this  kind  the  simplest  rule  can  be  used. 

179.  lu  Fig.  255  we  have  shown  the  method  of  finding  the  thrust  of  the  connect- 
ing-rod for  any  other  position  of  the  crank.  Let  D  P  represent  the  center  line  of 
motion ;  D,  the  center  of  driving  axle ;  C,  the  center  of  crank-pin ;  D  C,  the  position  of 
the  center  line  of  crank ;  and  the  circumference  R  C,  the  path  of  the  center  C.  Now, 
in  order  to  find  the  thrust  of  the  connecting-rod  when  the  crank-pin  is  in  this  position, 


Fig.  355 


we  draw  through  the  center  C  a  line  A  C,  perpendicular  to  the  center  line  of  motion 
D  P.  And  from  C  as  a  center,  and  with  a  radius  equal  to  the  length  of  the  connect- 
ing-rod, describe  a  short  arc  cutting  the  line  D  P  at  the  point  P ;  join  C  and  P  by  a 
straight  line,  thus  completing  the  triangle  A  C  P,  shown  in  heavy  lines  for  distinction. 
The  line  A  P  will  again  represent  the  total  steam  pressure  on  the  piston,  and  the  line 
A  C  the  thrust,  or  the  pressure  of  the  crosshead  against  the  guide.  To  find  the 
magnitude  of  the  force  represented  by  the  line  A  C,  we  have  the  following  approxi- 
mate rule : 

RULE  22. — Divide  the  length  of  the  line  A  C  in  inches  (that  is,  the  perpendicular 
distance  from  the  center  of  the  crank-pin  to  the  center  line  of  motion)  by  the  length 
of  the  connecting-rod  in  inches,  and  multiply  the  quotient  by  the  total  pressure  of 
steam  on  the  piston ;  the  product  will  be  the  thrust  in  pounds. 

Thus:  EXAMPLE  46. — If  the  total  pressure  is  24,120  pounds,  the  length  of  the 
connecting-rod  is  84  inches,  and  the  length  of  the  line  A  C  is  equal  to  8  inches  (by 
measurement),  what  will  be  the  thrust  of  the  connecting-rod  ? 

/T  x  24,120  =  2,287  pounds  =  pressure  of  the  crosshead  against  the  guides,  or  thrust  of  the  connecting-rod. 

Should  it  be  desirable  to  obtain  this  thrust  more  accurately,  then  apply  Rule  19, 
and  instead  of  using  the  length  of  the  crank  as  stated  in  that  rule,  use  the  length  of 
the  line  A  C  in  Fig.  255 ;  the  length  of  the  line  A  C  is  always  found  by  measurement. 
Remember  that  in  all  these  examples  the  steam  follows  full  stroke. 

180.  In  passenger  locomotives,  or  other  locomotives  in  which  the  crank  generally 
turns  in  the  same  direction,  the  upper  crosshead  gibs  will  wear  faster  than  the  lower 
gibs.  This  is  explained  in  the  following  manner : 

Let  the  arrow  2  in  Fig.  254A  represent  the  direction  in  which  the  crank  turns 
when  the  locomotive  is  running  in  a  forward  direction ;  then  when  the  crank-pin  is  at 
C2  the  reaction  of  the  connecting-rod  will  be  in  the  direction  of  the  arrow  3,  causing 
the  crosshead  to  be  forced  against  the  upper  guide,  as  indicated  by  arrow  5 ;  when  the 
crank-pin  is  at  (73  the  reaction  of  the  connecting-rod  will  be  in  the  direction  of  arrow 
4,  and  again  force  the  crosshead  against  the  guides,  as  indicated  by  arrow  5  as  before. 


MODEIty  LOCOMOTIVE  COXSTRUCTWX.  163 

In  fact,  throughout  the  stroke  when  the  engine  is  going  ahead,  the  erosshead  will 
press  against  the  upper  guide;  but  it  should  be  remembered  that,  when  the  center  line 
of  crank  stands  perpendicular  to  the  center  line  of  motion,  then  the  pressure  of  the 
crosshcud  against  the  guides  will  be  the  greatest;  or,  we  should  say,  that  in  the  neigh- 
borhood of  the  center  of  the  guides  the  pressure  will  be  the  greatest,  and  will  gradually 
decrease  as  the  crosshead  approaches  the  ends  of  the  guides.  In  switching  engines 
which  are  run  as  often  backward  as  forward,  the  upper  and  lower  gibs  of  the  cross- 
head  will  wear  very  nearly  alike,  because  in  running  forward  the  pressure  of  the 
erosshead  is  against  the  upper  guides,  and  in  running  backward  the  pressure  is  against 
the  lower  guides. 

PROPORTIONS  OF  CROSSHEAD. 

181.  The  sliding  surfaces  of  a  erosshead  should  not  be  too  large,  as  this  will 
make  the  erosshead  too  heavy ;  neither  should  they  be  too  small,  because  with  small 
sliding  surface  the  pressure  per  square  inch  on  these  surfaces  will  be  increased  to 
an  extent  which  will  heat  the  guides  and  cause  abrasion  or  cutting.  Now  it  will 
be  apparent  that  before  we  can  determine  the  dimensions  of  the  sliding  surfaces,  we 
must  know  the  pressure  which  can  be  allowed  per  square  inch  on  these  surfaces  for  the 
best  practical  results.  Knowing  this  pressure,  and  also  the  total  pressure  of  the  cross- 
head  against  the  guides,  which  must  be  provided  for,  the  calculations  for  obtaining  the 
dimension  of  the  sliding  surfaces  or  the  length  and  breadth  of  the  gibs  will  be  an  easy 
matter. 

According  to  the  dimensions  of  crossheads  in  different  classes  of  locomotives 
made  by  various  builders  and  master-mechanics,  and  under  the  assumption  that  the 
maximum  steam  pressure  in  the  cylinders  is  120  pounds  per  square  inch,  the  writer 
finds  that  50  pounds  per  square  inch  of  sliding  surface  is  a  good  average;  in  a 
few  cases  the  pressure  per  square  inch  was  somewhat  less  than  50  pounds,  and  in  a 
number  of  crossheads  75  pounds  per  square  inch  was  reached.  In  the  writer's  opinion 
50  pounds  pressure  per  square  inch  will  give  very  good  results,  and  may  be  adopted 
for  determining  the  dimensions  of  the  sliding  surface  or  the  gibs  of  a  erosshead  which 
is  to  be  designed. 

In  the  following  we  shall  adopt  50  pounds  per  square  inch  as  the  standard.  The 
total  pressure  of  the  erosshead  against  the  guides  is  obtained  by  Rule  20  or  21.  Our 
next  step  will  be  to  determine  the  area  of  the  sliding  surface  or  gibs. 

RULE  23. — Divide  the  total  pressure  of  the  erosshead  against  the  guides  by  50 ; 
the  quotient  will  be  the  area  in  square  inches  of  the  gibs  or  sliding  surface. 

EXAMPLE  47. — The  cylinders  in  a  locomotive  are  17  inches  in  diameter;  stroke,  24 
inches;  steam  pressure,  120  pounds  per  square  inch;  length  of  connecting-rod,  7  feet: 
it  is  required  to  find  the  area  in  square  inches  of  the  erosshead  gibs. 

The  area  of  a  piston  17  indies  in  diameter  is  22<>.!>S  square  indies;  the  total 
pressure  on  the  piston  will  be  equal  to  226.98  x  120  =  272)17. l»  pounds.  According  to 
Rule  21,  the  total  pressure  of  the  erosshead  against  the  slides  will  be  equal  to  27237.6 
x  i  =  3891  pounds.  And  lastly,  according  to  Rule  23,  we  have  --J};!1-  =  77.8  square 
indies  for  the  area  of  the  gibs,  or  sliding  surface  of  Die  crosshcad.  The  dimensions  of 
erosshead  given  in  Fig.  240  are  those  of  a  erosshead  in  actual  use  for  a  17  x  24  engine, 


MODERN  LOCOMOTIVE  CONSTRUCTION. 

and  has  given  good  satisfaction.  In  this  crosshead  we  find  the  area  of  the  sliding 
surface  to  be  equal  to  (3"  +  3")  x  14"  =  84  square  inches.  According  to  our  rule,  it 
should  have  77.8  square  inches.  Here,  then,  is  a  difference  of  6.2  square  inches,  which 
is  due  to  the  fact  that  we  have  made  no  allowance  for  the  play  of  the  driving  box  in 
the  pedestal ;  this  play  should  have  been  added  to  the  length  of  crank ;  but  when  we 
remember  that  this  size  of  sliding  surface  is  sometimes  used  in  locomotives  having 
cylinders  18  inches  in  diameter,  we  may  conclude  that  the  area  found  according  to 
our  rule  will  give  satisfactory  results. 

Assuming  that  in  all  locomotives  the  maximum  steam  pressure  in  the  cylinder  is 
120  pounds  per  square  inch,  and  the  ratio  of  the  length  of  crank  to  length  of  the 
connecting-rod  is  as  1  to  7,  and  that  the  pressure  per  square  inch  of  the  crosshead 
sliding  surface  should  be  about  50  pounds,  we  may  find  the  area  of  the  sliding  surface 
by  the  following  simple  rule : 

RULE  24. — Multiply  the  area  of  the  piston  by  the  decimal  .34 ;  the  product  will  be 
the  area  in  square  inches  of  the  sliding  surface  or  of  gibs. 

This  rule  will  give  results  approximately  correct,  for  all  ordinary  locomotive  prac- 
tice in  which  the  maximum  pressure  does  not  exceed  120  pounds. 

EXAMPLE  48. — What  must  be  the  area  of  the  sliding  surface  of  a  crosshead  for  a 
locomotive  having  cylinders  20  inches  in  diameter  ? 

The  area  of  a  20-inch  piston  is  314.16  square  inches;  hence  314.16  x  .34  = 
106.81+  square  inches,  which  is  the  area  of  the  sliding  surface  of  the  crosshead.  Com- 
paring this  area  with  that  of  the  crosshead  shown  in  Fig.  244,  we  find  the  area 
obtained  by  calculation  to  be  slightly  in  excess  of  that  in  the  illustration. 

TO   FIND   THE   LENGTH   AND   BREADTH   OF   THE   GIB. 

182.  Examining  the  illustrations  of  the  crossheads,  we  find  that  the  length  of  each 
gib  is  very  nearly  equal  to  five  times  its  breadth;  in  some  it  is  more,  in  others 
a  little  less ;  let  us  adopt  the  ratio  of  1  to  5  as  the  correct  proportion  of  the  breadth 
to  the  length  of  these  surfaces.  Now,  when  we  know  the  area  of  the  gib  and  the 
ratio  of  its  length  to  breadth,  the  dimensions  of  the  latter  are  easily  obtained  by 
the  following  rule : 

RULE  25. — Divide  the  area  of  the  surface  by  5,  and  extract  the  square  root  of  the 
product ;  the  answer  will  be  the  breadth  in  inches ;  this  breadth  multiplied  by  5  will 
give  the  length. 

EXAMPLE  49. — The  area  of  a  gib  or  sliding  surface  is  45  square  inches:  it  is 
required  to  find  the  length  and  breadth  of  the  gib. 

45  -T-  5  =  9,  and  the  square  root  of  9  or  V9  =  3  inches. 

Three  inches  is  the  breadth,  and  3  x  5  =  15  inches  is  the  length  of  the  gib. 

EXAMPLE  50. — What  should  be  the  length  and  breadth  of  a  gib  for  a  crosshead 
with  two  guides  in  a  locomotive  having  cylinders  18  inches  in  diameter  ? 

The  area  of  a  piston  18  inches  in  diameter  is  254.47  square  inches. 

The  area  of  the  gib  or  sliding  surface,  according  to  Rule  24,  must  be  254.47  x 
.34  =  86.5198  square  inches.  Then  by  Rule  25  we  have  86.5198  4-  5  =  17.3039 


MODEHX   LOCOMOTIVE    CONSTRUCTION. 


165 


and  ^17.3039  =  4.15  inches  for  the  breadth  of  the  gib ;  and  4.15  x  5  =  20.75  inches 
for  the  length. 

EXAMPLE  51. — What  should  be  the  length  and  breadth  of  a  gib  for  a  crosshead 
with  four  guides  in  a  locomotive  having  cylinders  18  inches  in  diameter  ! 

In  the  solution  of  this  problem  it  should  be  remembered  that  when  a  crosshead 
has  four  guides,  as  shown  in  Figs.  234  and  237,  the  area  of  the  sliding  surface  is  the 
sum  of  the  areas  of  two  gibs,  as  F  and  G  in  Fig.  257;  hence,  notice  the  following 
solution : 

The  total  area  of  the  sliding  surface  will  be  254.47  square  inches  x  .34  =  86.5198 
square  inches,  as  in  Example  50 ;  but,  as  already  stated,  this  sliding  surface  is  made 
up  of  two  gibs,  hence  the  area  of  each  gib  must  be  86.5198  -j-  2  =  43.2599  square 
inches,  then  43.2599  -=-  5  =  8.6519,  and  V8.6519  =  2.94  inches  for  the  width  of  gib, 
say  3  inches ;  and  3  x  5  =  15  inches  for  the  length  of  the  gib. 

Gibs  are  generally  made  from  £  inch  to  §  inch  thick. 


WIDTH   OF   CKOSSHEADS   AND   DIAMETER   OF   HUBS. 

183.  In  locomotive  cast-iron  crossheads  into  which  the  piston  rods  are  fitted  with 
tapered  enils,  and  sometimes  not  resting  against  a  shoulder,  as  shown  in  Fig.  255A,  the 
area  of  the  hub  around  the  key,  as  shown  in  Fig.  256,  should  be  sufficiently  large,  so 


fiff.  2SS.I 


Fig.  HS0 


froth. 


«-~  a-— 


Fig.  257 


that  the  tensile  stress  or  pull  will  not  exceed  3,000  pounds  per  square  inch  of 
section.  This  section  is,  of  course,  assumed  to  be  taken  through  D  E,  Fig.  255A, 
where  the  area  of  the  hub  will  be  smallest.  Hence,  to  find  the  diameter  B  of  the  hub, 
Fig.  '256,  or  the  width  A,  Fig.  12.">7,  wo  may  use  the  following  approximate  rule: 

RULE  26. — For  cast-iron  crossheads,  multiply  the  large  diameter  of  the  tapered  end 
of  piston-rod,  as  given  in  Table  14,  by  2 ;  the  product  will  be  the  outside  diameter 
of  the  hub. 

"When  the  piston-rod  has  no  shoulder  fitting  ;ig;iinst  the  end  of  the  hub,  as  shown 
in  Fig.  255A,  then  make  the  outside  diameter  of  the  hub  equal  to  twice  the  diameter  of 
the  tapered  end  measured  at  C — that  is,  the  diameter  of  the  hole  in  the  face  of  the  hub. 

For  wrought-iron  crossheads  multiply  the  large  diameter  of  the  tapered  end,  or 
the  diameter  at  C,  by  1.8 ;  the  product  will  be  the  outside  diameter  B  of  the  hub,  Fig. 
256,  or  the  width  A,  Fig.  257. 


1GG  MODERN  LOCOMOTIVE   CONSTRUCTION. 

EXAMPLE  52. — Find  the  diameter  or  the  width  of  a  locomotive  cast-iron  crosshead 
suitable  for  four  guides,  cylinders  17  inches  in  diameter. 

According  to  Table  14,  the  large  diameter  of  the  tapered  end  of  a  piston-rod  for  a 
cylinder  17  inches  in  diameter  is  2f  inches ;  hence  2f  x  2  =  5 J  inches  for  the  outside 
diameter  of  hub. 

EXAMPLE  53. — What  must  be  the  diameter  of  the  hub  of  a  wrought-iron  crosshead 
suitable  for  two  guides,  cylinders  20  inches  in  diameter?  According  to  Table  14,  the 
large  diameter  of  the  tapered  end  of  the  piston-rod  is  3  inches;  hence  3  x  1.8  =  5.4 
inches,  which  is  the  diameter  of  hub. 

The  width  of  the  latter  class  of  crossheads  will  frequently  have  to  be  determined 
by  the  width  of  guides,  or  the  length  of  the  crosshead  pin. 

DIMENSIONS    OP    GUIDES. 

184.  The  guides  should  be  made  as  short  as  possible.  In  practice  half  an  inch 
for  clearance  at  each  end  of  the  crosshead  is  generally  considered  to  be  the  least 
amount  that  should  be  allowed.  Consequently,  if  in  Fig.  234  the  stroke  is  24  inches, 
the  length  of  the  sliding  surfaces  (which  in  this  case  is  equal  to  the  length  of  the  gibs) 
is  equal  to  15$  inches,  and  the  clearance  at  each  end  is  equal  to  half  an  inch,  then 
the  shortest  distance  between  the  guide-blocks  is  equal  to  24"  +  15f "  +  J"  +  £"  =  40^ 
inches.  If  to  this  distance  is  added  the  necessary  amount  for  bolting  the  guide  to  the 
guide-blocks,  then  the  shortest  length  of  guides  will  have  been  obtained. 

Sometimes  the  general  design  of  a  locomotive,  generally  the  position  of  the 
driving  wheels,  will  compel  us  to  make  the  length  of  the  guides  greater  than  that 
determined  by  the  foregoing  figures. 

The  breadth  of  the  guide  is  equal  to  the  breadth  of  gib  or  sliding  surfaces  of  the 
crosshead,  according  to  rules  already  given. 

In  order  to  determine  the  thickness  of  a  guide  we  must  consider  it  to  be  a  beam 
firmly  fastened  at  the  ends  and  loaded  in  the  middle.  The  rule  for  finding  the  thick- 
ness of  a  beam  when  all  its  other  dimensions  and  load  are  known  is  as  follows: 
Multiply  the  length  of  beam  in  feet  by  the  load,  and  divide  this  product  by  the 
breadth  in  inches  multiplied  by  a  constant  number.  The  square  root  of  the  quotient 
will  be  the  thickness  in  inches.  The  constant  number  alluded  to  is  determined  by 
experiment,  and  is  not  the  same  for  different  kinds  of  material.  From  these  remarks 
we  would  infer,  and  correctly  too,  that  a  different  constant  should  be  used  for  steel 
than  is  used  for  iron  beams.  But  in  locomotive  practice,  for  the  sake  of  interchange- 
ability,  it  is  customary  to  use  the  same  constant  for  wrought-iron  and  steel  guides,  and 
this  practice  we  shall  follow  in  these  articles.  The  constant  number  used  in  calculat- 
ing the  strength  of  guides  is  1,200. 

In  these  calculations  we  shall,  for  the  sake  of  simplicity,  call  the  distance  between 
the  guide-blocks  the  "  length  of  the  guides."  Hence,  for  finding  the  thickness  of  a 
wrought-iron  or  steel  guide  we  have  the  following  rule : 

EULE  27. — Multiply  the  length  of  the  guide  in  feet  by  the  load  in  pounds ;  divide 
this  product  by  the  breadth  in  inches  into  the  constant  number  1,200 ;  extract  the 
square  root  of  the  quotient,  then  the  answer,  increased  by  an  amount  deemed  neces- 
sary for  re-planing  and  wear,  will  be  the  thickness  of  the  guide  in  inches. 


JfODffR.V  LOCOMOTIVE   CONSTRUCTION.  167 

EXAMPLE  54. — Find  the  thickness  of  the  guides  (two  guides  being  used  for  one 
crosshead)  for  a  locomotive  having  cylinders  20  inches  in  diameter,  24  inches  stroke, 
length  of  guides  4  feet,  breadth  of  guides  43  inches,  length  of  connecting-rod  7  feet, 
steam  pressure  in  cylinders  120  pounds  per  square  inch. 

In  the  first  place  we  must  find  the  load  which  these  guides  will  have  to  support. 
The  load  is  equal  to  the  greatest  pressure  of  the  crosshead  against  the  guide,  and 
consequently  can  be  found  by  Rule  21.  Now,  we  have  for  the  total  steam  pressure  on 
the  piston,  314.16  square  inches  x  120  pounds  =  37G99.2  pounds,  and  according  to 
Kule  21,  we  have  37699.2  x  |  =  5385.6  pounds  of  pressure  of  crosshead  against  the 
guides,  which  is  now  considered  to  be  the  load  that  one  guide  will  have  to  support. 

Now,  to  find  the  thickness  we  have,  according  to  Eule  27, 

=  3.77;  and  the  square  root  of  3.77,  or  V3J7  =  1.9,  say  2  inches; 
4.  t  D   X    j.wUO 

adding  to  this  J  of  an  inch  for  truing  up,  when  that  becomes  necessary  through  wear, 
we  have  2J  inches  for  the  thickness  of  guide. 

EXAMPLE  55. — Find  the  thickness  of  the  guides  (four  guides  being  used  for  one 
crosshead)  for  a  locomotive  having  cylinders  17  inches  in  diameter,  24  inches  stroke, 
length  of  guides  3  feet  4£  inches  (=  3.375  feet),  breadth  of  guide  3  inches,  length  of 
connecting-rod  7  feet,  and  steam  pressure  in  cylinder  120  pounds  per  square  inch. 

Area  of  piston  =  226.98 ;  hence  226.98  square  inches  x  120  =  27237.6  pounds  total 
steam  pressure  on  the  piston,  and  27237.6  x  |  =  3891.08  pounds,  which  is  the  load 
two  guides  bearing  against  the  gibs  F  and  G,  Fig.  257,  will  have  to  support ;  conse- 
quently one  guide  will  have  to  support 

3891.08 
—n- —  =  1945.54  pounds, 

and 

1945.54  x  3.375  _ 
3  x  1200  L8w' 

the  square  root  of  1.82,  or  ^1.82  =  1.34,  say  If  inches ;  add  to  this  i  of  an  inch  for 
truing  up,  we  have  1§  inches  for  the  thickness  of  guides. 

185.  Sometimes  the  thickness  of  the  guides  is  greater  at  the  center  of  their  length 
than  at  the  ends,  as  shown  in  Fig.  241 ;  this  is  done  to  reduce  the  weight  of  the  guides ; 
or,  in  other  words,  this  form  of  guide  will  give  the  maximum  strength  with  the  mini- 
mum amount  of  metal.     Tapering  the  guides  adds  considerably  to  the  expense  of  labor ; 
and  since  the  extra  weight  of  guides  with  parallel  faces,  or  of  straight  guides,  is  not 
objectionable,  the  writer  believes  that  the  latter,  such  as  are  shown  in  Fig.  260,  and 
whose  thickness  or  depth  throughout  is  equal  to  the  thickness  at  the  center  of  tapered 
guides  having  the  same  work  to  do,  are  more  desirable  to  use,  and  will  give  as  good,  if 
not  better,  results ;  again,  observing  that  the  majority  of  locomotives  have  straight 
guides,  we  are  led  to  believe  that  many  master-mechanics  and  locomotive  builders  share 
our  opinion. 

186.  In  determining  the  depth  of  a  guide  for  a  crosshead  requiring  one  guide  only, 
as  shown  in  Fig.  '24X,  good  judgment  must  )><•  used,  Ix-raiise  in  this  case  we  must  not 
only  make  the  guide  sufficiently  strong  to  resist  the  thrust  of  the  connecting-rod,  but 


1G8  MODERN  LOCOMOTIVE   CONSTRUCTION. 

we  must  also  make  it  deep  enough  to  guide  the  end  of  the  piston-rod  in  a  straight  line 
laterally.  It  should  also  be  remembered  that,  since  guides  of  this  kind  will  wear  on  the 
upper  and  lower  surfaces,  we  must  allow  double  the  amount  for  re-planing  that  has 
been  allowed  in  the  other  classes  of  guides ;  that  is  to  say,  when  only  one  guide  is  used 
for  a  crosshead,  then  we  should  allow  for  re-planing  £  of  an  inch  on  the  upper  sliding 
surface,  and  £  of  an  inch  for  the  lower  sliding  surface.  Taking  these  things  into 
consideration,  we  may  employ  for  determining  the  depth  or  thickness  of  a  guide  such 
as  shown  in  Fig.  248  the  following  rule  : 

RULE  28. — Find  the  depth  of  the  guide  by  Eule  27 ;  add  the  necessary  amount  for 
re-planing,  and  increase  this  sum  by  25  to  35  per  cent,  to  obtain  sufficient  depth  of 
guide  to  keep  the  piston-rod  in  a  straight  line  latei-ally. 

When  the  crank  is  turning  in  the  direction  of  arrow  1,  Fig.  260 — that  is,  when  the 
engine  is  running  ahead — the  weight  of  the  crosshead  has  a  tendency  to  reduce  the 
thrust  of  the  connecting-rod  against  the  upper  guide ;  but  when  the  engine  is  running 
in  the  direction  of  arrow  2 — that  is,  running  backward — the  thrust  of  the  connecting- 
rod  against  the  lower  guide  will  be  increased  by  the  weight  of  the  crosshead.  But  in 
locomotives  the  thickness  of  the  upper  guide  is  always  equal  to  that  of  the  lower  one. 
For  extreme  accuracy  in  determining  the  thickness  of  either  guide,  the  weight  of  the 
crosshead  should  be  taken  into  account;  and  furthermore  a  part  of  the  weight  of 
the  guide  should  also  enter  into  our  calculation.  For  the  sake  of  simplicity,  we 
omitted  referring  to  these  items  in  Eule  27,  but  they  were  taken  into  consideration 
when  the  constant  number  1,200  was  determined  upon.  Hence  these  rules  should 
only  be  used  for  determining  the  thickness  or  depth  of  locomotive  guides,  or  those  for 
engines  of  similar  design. 

CASTIB-ON  GUIDES. 

187.  Sometimes  locomotive  guides  are  made  of  cast-iron.     In  Figs.  258  and  259 
we  have  shown  the  form  and  given  the  dimensions  of  cast-iron  guides  working  satis- 
factorily in  ten-wheeled  engines  having  19  x  24  inch  cylinders. 

GUIDE   BOLTS. 

188.  The  guides  are  held  in  position  at  one  end  by  the  guide-yoke  y,  Fig.  260,  and 
at  the  other  end  they  are  fastened  to  the  cylinder  head.     The  guide-yoke  is  made 
strong  enough  to  prevent  the  end  of  the  guide  moving  in  a  vertical  direction,  but  has 
not  sufficient  strength  to  resist  a  force  acting  against  it  in  a  horizontal  direction, 
consequently  it  cannot  resist  the  horizontal  force  due  to  the  friction  between  the 
guides  and  crosshead.     Therefore,  for  holding  the  guides  in  position  only  one  bolt  is 
used  at  the  yoke  end,  and  two  bolts  are  used  at  the  cylinder  end.     The  bolt  at  the 
yoke  end  and  one  of  the  bolts  at  the  cylinder  end,  we  may  say,  are  for  the  purpose  of 
resisting  the  thrust  of  the  connecting-rod,  and  the  second  bolt  at  the  cylinder  end  is 
for  the  purpose  of  resisting  the  force  acting  in  a  horizontal  direction  or  the  pull  due  to 
friction  between  the  guides  and  crosshead.     These  bolts  are  generally  £  of  an  inch  in 
diameter  for  small  locomotives;    I  and   sometimes   1  inch   in   diameter  for  larger 
locomotives ;  and  occasionally  we  meet  with  bolts  l-^-  inches  in  diameter  for  locoino- 


MODER\  LOCOMOTIVE   CONSTRUCTION. 


169 


tives  having  cylinders  20  inches  in  diameter.  It  is  always  best  to  use  tapered  bolts 
for  this  purpose,  because  these  must  have  a  good  fit,  and  if  they  are  made  straight,  it 
would  require  too  much  labor  or  the  bolt  may  be  ruined  in  driving  it  out,  when  that 
becomes  necessary  to  take  out  the  thin  strips  of  copper  placed  between  the  guide-blocks 
and  guides  to  take  up  the  wear.  It  should  also  be  noticed  that  in  locomotives  having 
cylinders  of  the  same  diameters  but  different  design  of  crosshead,  the  diameters  of  the 
guide  bolts  in  the  designs  such  us  shown  in  Figs.  234  and  237  are  about  the  same  as 
those  used  in  the  design  shown  in  Fig.  2GO;  yet,  when  we  compare  the  manner  of 
fastening  the  guides  at  the  yoke  end  in  these  designs,  it  would  seem  that  in  the 


'-.* 


Fig.  259 


design  Fig.  260  guide  bolts  of  smaller  diameter  can  be  used  than  in  the  designs 
shown  in  Figs.  234,  237;  because  in  Hie  latter  two  designs  some  of  the  thrust  of  the 
connecting-rod  acts  directly  against  the  bolts ;  Avhereas  in  the  former  design  the  same 
amount  of  thrust  seems  1<>  act  directly  against  the  guide-blocks  B  B.  But  a,  closer 
examination  of  the  design  in  Fig.  260  will  make  it  apparent  that,  should  the  guides 
spring,  which  may  occur,  and  often  does  occur,  then  the  thrust  will  act  with  a  long 
leverage  against  the  bolt  which  has  but  a  very  short  leverage — the  edge  /  of  the  guide- 
block  being  the  fulcrum — and  thus  greatly  increase  the  stress  in  the  bolt  at  the  yoke  end. 
For  these  reasons,  then,  we  may  assume  that  the  amount  of  stress  in  the  guide  bolts  in 
design  Fig.  260  is  the  same  as  the  amount  of  stress  in  the  guide  bolts  in  designs  Figs. 
ii:i4  and  237  when  the  thrust  of  the  connecting-rod  in  one  case  is  equal  to  that  in 
the  other  case. 


170 


MODERN  LOCOMOTIVE    COXSTRVCTIOX. 


DISTANCE   BETWEEN   THE   GUIDES. 

189.  The  distance  from  a  to  &  between 
the  guides  in  Fig.  260  (this  distance  also 
determines  the  depth  of  the  crosshead) 
must  be  sufficient  to  clear  the  connect- 
ing-rod in  any  oblique  position  which 
it  may  occupy  during  the  revolution  of 
the  crank.  Our  first  step,  then,  will  be  to 
find  the  position  of  the  guide-yoke  y  when 
the  length  of  the  connecting-rod,  stroke, 
and  length  of  crosshead  are  given ;  and 
secondly,  we  must  find  that  oblique  posi- 
tion of  the  connecting-rod  which  will  re- 
quire the  greatest  distance  between  the 
guides.  Let  L  M  (Fig.  260)  represent  the 
center  line  of  motion  of  the  crosshead  pin, 
and  the  point  x  on  this  line  the  center  of 
the  axle.  From  the  point  £  as  a  center, 
and  with  a  radius  equal  to  half  the  stroke, 
describe  a  circle  5  s.2  s3 ;  the  circumference 
of  this  circle  will  represent  the  path  of  the 
center  of  the  crank-pin  when  the  center  of 
the  axle  is  at  x.  From  the  point  s,  at 
which  the  circumference  cuts  the  line  L  M, 
lay  off  on  the  center  line  of-  motion  L  M 
a  point  d ;  the  distance  between  the  points 
s  and  d  must  be  equal  to  the  length  of 
the  connecting-rod.  From  the  point  d  to- 
wards s  lay  off  a  point  » ;  the  distance 
between  the  points  d  and  H,  measured  on 
the  line  L  M,  must  be  equal  to  the  sum 
of  the  clearance  at  the  rear  end,  and  the 
horizontal  distance  between  the  center  of 
crosshead  pin  and  the  rear  end  of  the 
sliding  surfaces  of  the  crosshead,  or  the 
rear  end  of  the  gib  when  one  is  used; 

\  also,  on  the  line  X  M  lay  off  from  the  point 
«  towards  s  a  point  v ;  the  distance  be- 
tween the  points  n  and  v  must  be  equal  to 

A<  the  thickness  of  the  guide-yoke.  Through 
the  points  u  and  v  draw  straight  lines 
perpendicular  to  the  center  line  of  motion ; 
these  lines  will  represent  the  inner  and 
outer  faces  of  the  guide-yoke,  and  thus 


MODERX  LOCOMOTIVE  OOlTSTBVCTIOJr. 

establish  its  position.  This  position  of  the  guide-yoke  is  the  closest  to  the 
cylinder  head  which  it  can  occupy;  sometimes  the  guide-yoke  may  have  to  be  moved 
towards  the  center  of  axle  in  order  to  clear  the  driving  wheel  or  to  suit  other  parts  of 
the  design.  Secondly,  to  find  the  oblique  position  of  the  connecting-rod  which  will 
require  the  greatest  distance  between  the  guides.  Through  the  point  x  draw  a  straight 
line  0  P  perpendicular  to  L  M;  on  this  line  0  P  lay  off  above  the  center  x  a  point  #2; 
the  distance  between  the  points  x  and  jr.,  must  be  equal  to  the  distance  through  which 
the  driving  box  can  move  before  it  strikes  the  frame ;  hence  the  point  x2  will  be  the 
•  •cuter  of  the  axle  when  the  driving  box  occupies  its  highest  position  in  the  pedestal 
of  the  frame.  Again,  on  the  line  0  P  lay  off  below  the  center  x  a  point  x3 ;  the  distance 
between  the  points  x  and  x3  must  be  equal  to  the  distance  through  which  the  driving 
Itox  <-an  move  before  it  strikes  the  pedestal  cap;  hence  the  point  x3  will  be  the  center 
of  the  axle  when  the  driving  box  occupies  its  lowest  position  in  the  pedestal.  From 
the  point  x.2  as  a  center,  and  with  the  radius  equal  to  half  the  stroke,  describe  an  arc 
ef;  also,  from  the  point  x3  as  a  center,  and  with  the  same  radius,  describe  an  arc  .9  It. 
When  the  driving  box  is  in  its  extreme  highest  position  the  arc  e  /will  be  a  part  of 
the  path  of  the  center  of  crank-pin ;  and  when  the  driving  box  is  in  its  extreme  lowest 
position  the  arc  //  //  will  be  a  part  of  the  path  of  the  center  of  crank-pin.  From  the 
point  c2  at  which  the  line  0  P  cuts  the  arc  ef  lay  off  on  this  arc  a  number  of  points 
c3,  c4,  c5,  CG,  c- ;  the  distance  between  these  points  should  be  about  2  inches  or  a  little 
less.  From  these  points  as  centers,  and  with  a  radius  equal  to  the  length  of  the  con- 
necting-rod, describe  a  number  of  short  arcs  cutting  the  center  line  of  motion  L  M  at 
the  points  d3,  rf,,  d7l,  rf,;,  <!-.  Join  by  straight  lines  the  points  c3  and  (13,  c^  and  dv  c5  and 
d5,  etc. ;  then  the  line  c3  d3  will  represent  the  center  line  of  the  connecting-rod  when  the 
crank-pin  is  at  c3,  and  the  lines  ct  d±,  c5  d5,  etc.,  will  represent  the  center  line  of  con- 
necting-rod when  the  crank-pin  is  at  c4,  c3,  etc.  Now  select  that  center  line  of  the 
connecting-rod  which  cuts  the  inner  face  i  k  of  the  guide-yoke  at  a  point  furthest  from 
the  line  L  M.  In  our  illustration  the  line  c5  d5  cuts  the  line  i  k  at  the  point  /,  and 
there  is  no  other  line  representing  the  center  line  of  connecting-rod  which  will  cut  the 
line  i  k  at  a  point  above  the  point  7 ;  hence  the  line  c5  rfr,  will  represent  the  center  line 
of  the  connecting-rod  in  such  an  oblique  position  as  will  require  the  greatest  distance 
between  the  upper  guide  and  the  line  L  M,  and  consequently  the  upper  guide  must  be 
placed  high  enough  to  clear  the  connecting-rod  in  this  position.  Above  the  center 
line  f5  (15  draw  a  line  m  n  to  represent  the  upper  edge  of  the  connecting-rod ;  the  space 
between  the  line  c5  dj  and  the  line  m  n  must  be  equal  to  that  between  the  center  line 
of  connecting-rod  and  its  outer  edge  at  a  corresponding  distance  from  either  end  of 
the  rod  to  the  guide-yoke;  if  the  edges  of  the  rod  are  parallel  to  each  other,  then  of 
course  the  line  in  n  will  also  be  parallel  to  c:>  d-, ;  if  the  rod  is  tapered,  then  the  line  m  n 
must  have  the  same  inclination  to  c5  d5  as  the  edges  of  the  rod  have  to  its  center  line. 

The  line  m  n  cuts  the  line  I  k  at  the  point  r,  and  theoretically  the  distance  between 
the  points  u  and  r  would  be  the  required  distance  between  the  center  line  of  motion 
L  M  and  the  lower  face  of  the  upper  guide.  But,  since  small  inaccuracies  in  workman- 
ship are  very  difficult  to  avoid,  we  must  allow  some  extra  space  for  these  inaccuracies; 
and  besides  this  the  ends  of  the  slot  in  the  yoke,  shown  in  Fig.  261,  are  usually 
semicircular,  and  the  face  of  the  guide  is  placed  even  with  the  ends  of  the  slot,  or,  in 


172  MODERN  LOCOMOTIVE   CONSTRUCTION. 

other  words,  tangent  to  the  curved  part,  and  therefore  some  allowance  in  the  distance 
between  the  guides  must  be  made  to  enable  the  edges  of  the  rod  to  clear  the  curved 
surface  in  the  slot  of  the  yoke.     Therefore,  above  the  point  r  lay  off  on  the  line  i  k 
a  point  r2;   the  distance  between  the  points  r  and  r2  must  be 
equal  to  a  clearance  considered  to  be  sufficient  to  allow  for  the 
inaccuracy  of  workmanship  and  the  semicircular  form  of  the  end 
O  of  the  slot.     In  the  case  before  us  let  the  distance  between  the 

points  r  and  r2  be  equal  to  $  an  inch.  Through  the  point  r2 
draw  a  horizontal  line  r2  a ;  this  will  be  the  lower  surface  of  the 
upper  guide,  and  thus  the  position  of  the  upper  guide  is  estab- 
lished. Frequently  the  distance  between  the  centers  x  and  x2 
is  greater  than  the  distance  between  the  centers  x  and  #3; 
Fig.  261  hence  it  may  appear  that  the  distance  between  the  center  line  L  M 
and  the  upper  guide  should  be  greater  than  that  between  the  line 
L  M  and  the  lower  guide ;  but  in  such  cases,  after  the  position  of 
the  upper  guide  has  been  determined,  we  simply  place  the  lower 
one  at  a  distance  from  L  M  equal  to  that  from  L  M  to  the  upper 
guide. 

If,  on  the  other  hand,  the  distance  between  x  and  x3  is  greater 

than  from  x  to  x2,  then  we  must  find,  by  a  construction  similar  to  the  foregoing,  the 
distance  from  L  M  to  the  lower  guide,  and  place  the  upper  one  at  the  same  distance  so 
found  above  L  M,  because  it  is  always  desirable  to  have  the  guides  equidistant  from 
the  crosshead  pin. 

The  construction  in  Fig.  260  shows  plainly  that  the  distance  between  the  guides 
depends  upon  the  position  of  the  guide-yoke.  Again,  drawing  the  connecting-rod  in 
a  position  so  that  its  center  line  will  be  tangent  to  the  path  of  the  crank-pin,  as  shown 
by  the  line  c3  d3  (Fig.  260),  and  then  making  the  distance  between  guides  to  clear  only 
this  oblique  position  of  connecting-rod,  as  is  often  done,  may  result  in  bringing  the 
guides  too  close  together,  and  lead  to  considerable  trouble,  annoyance,  and  waste  of 
labor  in  having  to  chip  the  ends  of  the  guides,  and  increase  the  length  of  slot  in 
guide-yoke. 

PROPORTIONS  OF  CROSSHEAD  PINS. 

190.  In  establishing  rules  for  determining  the  dimensions  of  a  locomotive  cross- 
head  pin  it  will  be  best  to  base  these  rules  on  the  dimensions  of  pins  used  in  locomo- 
tives in  actual  and  successful  service.  A  great  difference  in  the  sizes  of  these  pins 
exists ;  but  in  the  writer's  opinion  the  sizes  given  in  Table  15  are  a  good  average,  and 
will  be  used  in  establishing  the  following  rules.  The  greater  number  of  the  smaller 
pins  whose  dimensions  are  here  given  are  made  of  cast-iron,  and  the  greater  number 
of  the  large  pins  are  made  of  wrought-iron. 


MODERX   LOCOMOTIVE    COXSTRrCTIOX. 


173 


TABLE   15. 

AVERAGE  DIMENSIONS  OF  CROSSHEAD  PINS  AS  AT  PRESENT  USED  IN  LOCOMOTIVES. 


Diameter  of 
Cylinder. 

Stroke. 

IlilllMftIT   Of 

Croeshead  Pin*. 

Length  or 
CroMhead  Plus. 

12" 

20" 

2" 

2" 

13" 

22" 

2*" 

3" 

14" 

22" 

2f" 

3" 

15" 

22" 

2*" 

3" 

15" 

24" 

3" 

3" 

16" 

22" 

3" 

3" 

16" 

24" 

3i" 

34" 

17" 

22" 

H" 

34" 

17" 

24" 

34" 

34" 

18" 

22" 

34" 

34" 

18" 

24" 

3i" 

•        3i" 

19" 

22" 

3*" 

3J" 

19" 

24" 

8t" 

34" 

20" 

24" 

»" 

3|" 

21" 

24" 

3f 

3f" 

N 

^- 

s 

!     fl— 

ft  —  ^"H 

j          | 

?      \ 
V      ; 

7? 

"<!* 
^ 

,     i 

c 

t| 

.      i          T 

JF 

>—  -< 

M- 

2"T 

^_                } 

i                       i 

^~ 

I 

Fig.  262 


Crosshead  pins  are  subjected  to  a  shearing  stress,  and  therefore  it  would  seem 
that  by  making  the  diameter  of  a  pin  sufficiently  large  to  give  it  the  necessary  strength 
to  resist  shearing,  and  then  assigning  to  it  some  given  length,  would  be  all  that  is 
required.  But  in  examining  the  dimensions  given  in  the  table  we  find  the  diameters 
of  these  pins  to  be  larger  than  is  necessary  for  their  adequate  strength.  Another 
notable  feature  is  that  the  length  of  the  pins  is,  in  all  cases  excepting  three,  equal 
to  the  diameter.  Here,  then,  we  conclude,  and  s»  _ 

rightly  too,  that  there  must  be  other  consid- 
erations besides  that  of  strength  which  the 
designers  have  kept  in  view  in  determining 
the  dimensions  of  these  pins. 

Let  us  first  turn  our  attention  to  the 
length  of  these  pins.  We  find  that  their 
lengths  compared  with  their  diameters  are 
less  than  the  lengths  of  pins  ordinarily 
used  in  stationary  or  marine  engines.  The 
reason  for  making  the  length  of  a  locomotive 
crosshead  pin  comparatively  so  short  is  that  considerable  lateral  play  is  allowed 
between  the  hubs  of  the  driving  wheels  and  their  boxes ;  and  consequently,  when  the 
locomotive  is  running  over  a  curve  the  mechanism  between  cylinders  and  wheels  will 
become  out  of  alignment,  which  will  create  an  extra  stress  on  the  crosshead  pin. 
Now,  in  order  to  reduce  this  extra  stress  as  much  as  possible,  the  pin  should  be  made 
as  short  as  good  practice  will  allow,  and  thus  reduce  the  leverage  of  the  pin.  Another 
ivason  for  making  these  pins  so  short  is  that  in  many  locomotives  the  space  for  the 
crosshead,  and  consequently  its  pin,  is  limited. 

From  these  considerations,  and  also  from  the  dimensions  given  in  the  table,  it  will 
be  seen  that  the  length  of  a  locomotive  crosshead  pin  should  be  equal  to  its  diameter. 

By  the  expression  "  length  of  crosshead  pin  "  is  meant  the  length  marked  L  in 
Fig.  262 — that  is,  the  length  of  that  part  of  the  pin  which  is  covered  by  the  connect- 


174  MODEMS  LOCOMOTIVE  CONSTRUCTION. 

ing-rod  brass ;  and  by  "  diameter  of  pin "  is  meant  the  diameter  marked  D — that  is, 
the  diameter  of  the  same  part  of  the  pin. 

Our  next  step  will  be  to  establish  a  rule  for  finding  the  diameter  of  the  pin. 
Here  a  consideration  of  the  greatest  importance  presents  itself,  namely,  we  must  not 
only  make  the  pin  large  enough  to  give  it  the  required  strength  to  do  the  work,  but  it 
must  also  have  a  large  working  surface,  so  as  to  avoid  heating.  Right  here  it  may  be 
remarked  that  when  locomotive  crosshead  pins  are  correctly  proportioned  so  that  they 
will  not  heat,  they  will  also  have  adequate  strength  for  the  work,  and  for  this 
reason  we  will  leave  the  consideration  of  strength  out  of  the  question.  But  it  must 
be  distinctly  understood  that  these  remarks  apply  only  to  locomotive  crosshead  pins, 
or  pins  which  are  as  short  in  comparison  with  their  diameters. 

The  pressure  on  the  piston  is  transmitted  to  the  crosshead  pin,  and  experience 
has  shown  that  when  the  pressure  on  the  crosshead  pin  exceeds  a  certain  amount  the 
oil  will  be  forced  out  of  the  bearing,  and  consequently  heating  or  abrasion  will  follow, 
and  the  only  means  at  hand  to  avoid  such  results  is  to  make  the  working  surfaces 
sufficiently  large  to  reduce  the  pressure  per  square  inch.  The  area  of  the  working 
surface  is  estimated  by  the  projected  area  as  shown  at  A  in  Fig.  262.  This  area  is 
always  equal  to  that  of  a  rectangle  whose  length  and  breadth  is  equal  to  the  length 
and  diameter  of  the  pin. 

Since  the  pressure  per  square  inch  on  the  pin  is  estimated  by  the  pressure  per 
square  inch  on  its  projected  area,  we  must  proportion  this  area  in  a  manner 
which  will  not  allow  the  pressure  to  exceed  a  given  limit.  Here,  then,  the  question 
arises,  what  is  this  limit?  To  answer  this  question  let  us  find  the  projected  area  of 
each  pin  given  in  Table  15,  which  is  obtained  by  multiplying  the  length  by  the 
diameter;  the  product  will  be  the  area  required.  Let  us  assume  that  the  greatest 
pressure  per  square  inch  on  the  piston  is  120  pounds ;  then  the  total  pressure  on  the 
piston,  and  therefore  on  the  projected  area  of  the  pin,  will  be  equal  to  the  product 
obtained  in  multiplying  the  area  of  the  piston  by  120. 

Now,  dividing  this  product  by  the  projected  area  of  the  pin  we  will  obtain  a 
quotient  which  will  be  the  pressure  per  square  inch.  By  so  doing  we  find  that  the 
pressure  per  square  inch  of  the  projected  area  varies  from  about  2,200  to  3,200 
pounds ;  and  since  these  pins  in  Table  15  have  given  satisfaction,  and  since  their 
dimensions  are  suitable  for  cast-iron  pins,  and  are  also  often  used  for  wrought- 
irou  pins,  we  may  adopt  either  2,200  or  3,200,  or  any  other  figure  between  these 
two,  as  a  limit  or  standard  of  pressure  per  square  inch  of  projected  area.  Let 
us  adopt  2,880  pounds  per  square  inch  as  a  standard.  We  are  now  in  a 
position  to  establish  a  rule  for  finding  the  dimensions  of  any  locomotive  crosshead 
pin. 

RULE  29. — Divide  the  total  pressure  on  the  piston  in  pounds  by  2,880  and  extract 
the  square  root  of  the  quotient ;  the  result  will  be  the  diameter  and  the  length  of  the 
pin  in  inches.  Or,  putting  this  rule  in  the  form  of  a  formula,  we  have, 

/Area  of  piston  in  sq.  inches  x  pressure  per  inch  of  piston   =  diameter  and  length  of  CI.OS8head  pin  in  inches. 

2880 

EXAMPLE  56. — Find  the  diameter  of  a  locomotive  crosshead  pin  suitable  for  a 


MODESX  LOCOMOTIVE   COXSTRTCTIOX. 


175 


cylinder  18  inches  in  diameter,  and  a  steam  pressure  of  120  pounds  per  square  inch  of 

piston. 

Area  of  18-inch  piston  =  254.47 ; 
hence, 

LT.4.47  x   li'O 

"2880"       :1°-60; 

mid  the  square  root  of  10.60  is  3J  (nearly) ;  therefore  the  diameter  of  the  pin  will  be 
3}  inches,  and  the  length  will  also  be  3J  inches. 

In  a  similar  manner  we  can  obtain  the  dimensions  of  the  crosshead  pins  for  all 
locomotives  when  the  diameter  of  cylinder  and  steam  pressure  is  given.  But  if  we 
assume  that  in  all  locomotives  the  maximum  steam  pressure  per  square  inch  of  piston 
is  120  pounds,  then  the  foregoing  rule  can  be  made  simpler,  and  we  obtain  the  following : 

RULE  30. — Divide  the  area  of  the  piston  in  square  inches  by  24,  and  extract  the 
square  root  of  the  quotient ;  the  answer  will  be  the  diameter  and  length  of  the  cross- 
head  pin. 

Or,  putting  this  rule  in  the  shape  of  a  formula,  we  have, 


'Area  of  piston  in  square  inches 
24 


=  diameter  and  length  of  crosshead  pin  in  inches. 


By  this  last  rule  the  dimensions  of  the  crosshead  pins  in  Table  16  have  been 
obtained.  As  will  be  seen,  the  dimensions  of  pins  in  Table  16  agree  very  closely  with 
those  in  Table  15,  and  may  therefore  be  adopted  as  the  standard  sizes  of  cast-iron 
pins  for  locomotives  in  which  the  steam  pressure  in  the  cylinders  does  not  exceed  120 
pounds  per  square  inch.  A  number  of  locomotive  builders  make  the  diameters  of 
wrought-iron  pins  somewhat  less  than  that  of  cast-iron  ones,  but  leave  the  length  of 
the  former  the  same  as  that  of  the  latter.  But,  on  the  other  hand,  quite  a  number  of 
builders  make  the  diameters  of  cast-iron  and  wrought-iron  pins  alike,  and  thereby 
obtain  a  greater  uniformity  in  the  patterns  for  the  connecting-rod  brasses. 


TABLE  16. 

DIMENSIONS   OF   CROSSHEAD   PINS   SUITABLE   FOR    LOCOMOTIVES   IN   WHICH  THE   STEAM 
PRESSURE  PER   SQUARE  INCH   DOES  NOT  EXCEED   120  POUNDS. 


Diameter  of  Cylinder*. 

Diameter  of  Croeshead  Pins. 

Length  of  Croeshead  Pint. 

9" 

If 

If 

10" 

«" 

If 

11" 

2" 

2" 

12" 

8j" 

2f 

13" 

2f 

2f 

14" 

2|" 

2f 

15" 

2f" 

2f" 

16" 

2S" 

2f 

17" 

3" 

3" 

18" 

3f 

3i" 

19" 

34" 

3f 

20" 

3f 

3f 

21" 

3i" 

3J" 

22" 

H" 

3f 

176 


MODERN  LOCOMOTIVE   COXSTUVCTION. 


REMARKS  RELATING  TO   THE  FORM   OF  CROSSHEAD  PINS. 

191.  Cast-iron  pins  are  cast  to  the  crosshead ;  wrought-iron  or  steel  pins  are  put 
in  separately,  and  fit  into  tapered  holes  in  the  crosshead.  The  taper  of  these  holes 
should  not  be  too  great,  as  an  excessive  taper  will  throw  too  much  stress  on  the  nuts 
and  is  liable  to  tear  the  pin  at  7?,  Fig.  262.  A  good  taper,  and  one  which  is  often  used, 
is  1J  inches  in  12  inches — that  is,  the  diameter  of  one  end  of  the  12  inches  will  be  l£ 


Fig.  263 


Fig.  264, 


inches  less  than  at  the  other  end.  Many  master-mechanics  use  one  taper  reamer 
only,  as  shown  in  Fig.  263,  for  reaming  both  holes  which  fit  the  pin  at  C  and  F. 
Some  master-mechanics  object  to  this  plan,  because  in  their  opinion  it  does  not  leave 
a  shoulder  sufficiently  large  at  E,  and  therefore  the  diameter  of  the  pin  at  C  is 
reduced,  although  the  taper  per  inch  is  left  the  same  as  before.  For  pins  of  this  kind 
a  step  reamer  similar  to  that  shown  in  Fig.  264  will  have  to  be  used. 

The  crosshead  pin  is  held  in  position  by  two  nuts  N  N,  and  to  obtain  further 
security  a  split  pin  P  is  inserted  in  the  end  of  crosshead  pin. 

All  crosshead  pins  should  be  prevented  from  turning  by  a  dowel  pin  or  a  small 
feather,  as  shown  at  C. 

Crosshead  pins  made  of  wrought-iron  should  be  case  hardened. 


STUFFING  BOXES  AND  GLANDS. 

192.  The  purpose  of  the  stuffing  box  and  gland  is  simply  to  hold  some  kind  of 
packing  close  against  the  piston-rods,  valve-rods,  or  spindles  of  valves,  etc.,  thus 
forming  a  steam-tight  passage  for  the  rods  or  spindles.  The  packing  may  be  divided 
into  two  distinct  classes,  namely,  metallic  packing  and  hemp  packing.  In  the  latter 
we  include  all  packing  made  of  fibrous  material. 

Most  of  the  metallic  packing  at  present  in  use  is  manufactured  by  firms  who 
make  the  manufacture  of  it  a  special  business,  and  these  firms  furnish  the  dimensions 
of  the  stuffing  boxes  to  hold  this  packing ;  hence  the  only  kind  of  stuffing  boxes 
to  be  considered  here  will  be  those  which  are  to  hold  hemp  packing. 

Fig.  265  represents  a  stuffing  box  with  gland  similar  to  those  cast  to  locomotive 
cylinder  heads.  Fig.  266  represents  a  stuffing  box  with  gland  similar  to  those  cast  to 
locomotive  steam  chests.  Fig.  267  represents  a  stuffing  box  with  screw-cap,  as  is 
generally  used  for  small  valves,  cocks,  etc. 

In  order  to  find  the  principal  dimensions  of  the  stuffing  box,  glands,  and  studs,  the 
following  rules  may  be  used ;  and  the  dimensions  thus  obtained  will  agree  with  good 
practice : 

RULE  31. — To  find  the  thickness  t  of  the  stuffing  box  shown  in  Figs.  265  and'  266, 


AIODEHX  LOCOMOTIVE   COSSTBVCTION. 


177 


Add  |  of  an  iuch  to  i  the  diameter  of  the  rod ;  the  sum  will  be  the  thickness  of 
the  stuffing  box  at  t. 

EXAMPLE  57. — Find  the  thickness  at  t  for  a  piston-rod  stuffing  box,  the  piston-rod 
being  24  inches  in  diameter. 

|  diam.  +  £  of  an  inch  =  thickness ; 
hence, 

I"  +  i"  =  £",  thickness  at  t. 

EXAMPLE  58. — What  should  be  the  thickness  at  t  for  a  valve-rod  stuffing  box,  the 
valve-rod  being  14  inches  in  diameter? 


hence, 


J  of  14.  inches  =  f ; 
I"  +  1"  =  f",  thickness  at  t. 


In  practice  the  thickness  at  t,  Fig.  266,  is  about  4.  inch  for  small  valve-rods, 
and  £  of  ail  inch  for  the  large  valve-rods.     The  thickness  at  t  for  piston-rod  stuffing 


boxes,  Fig.  265,  ranges  generally  from  3  inch  for  small  piston-rods  and  1  to  If  inches 
for  the  larger  ones. 

The  thickness  of  the  stuffing-box  flange  /  should  be  sufficient  to  obtain  a  good 
depth  of  thread  for  the  studs ;  hence,  to  find  thickness  /  of  the  stuffing-box  flange, 
we  have  the  following  rule : 

RULE  32. — To  i  of  the  diameter  of  the  rod  add  8  of  an  inch,  and  increase  this 
sum  25  per  cent.  This  sum  will  be  the  thickness  of  the  stuffing-box  flange.  Putting 
this  rule  in  the  shape  of  a  formula,  we  have, 


diameter  of  rod  4-  £  inch 


diameter  of  rod  +  |  inch 


==  thickness  of  the  stuffing-box  flange. 


EXAMPLE  59. — Find  the  thickness  of  a  valve-rod  stuffing-box  flange,  the    rod 
being  14  inches  in  diameter. 

J  of  14.  inches  =  |  inch ; 
hence, 

3"  +  8" 
I"  +  |"  +  •  — 1-=  -  =  U  inch  thickness  of  flange. 


178  MODERN  LOCOMOTIVE   CONSTRUCTION. 

The  flange  of  the  piston-rod  stuffing  box,  Fig.  265,  generally  forms  part  of  the 
cylinder  head  casing,  and  therefore  it  is  often  made  from  14  inches  to  16  inches  in 
diameter.  The  guide-blocks  are  also  fastened  to  this  flange,  and  consequently  its 
thickness  cannot  be  exactly  determined  by  the  foregoing  rule ;  the  thickness  of  the 
flange  near  the  body  of  the  box  will  always  be  more  than  that  obtained  by  this  rule. 

In  large  engines  the  thickness  of  this  flange  will  often  be  from  If  to  2  inches,  and 
this  thickness  extends  from  the  stuffing  box  to  beyond  the  guides;  and  from  the 
guides  to  the  edges  of  the  flange,  its  thickness  is  reduced  to  about  £  inch. 

193.  For  determining  the  packing  thickness  p  we  may  employ  the  following  rule : 
EULE  33. — To  J  of  the  diameter  of  the  rod  add  J  of  an  inch ;  the  sum  will  be  the 

thickness  p  of  the  packing.     Putting  this  rule  in  the  shape  of  a  formula,  we  have, 

i  diameter  of  rod  +  J  inch  =  thickness  p  of  packing. 

EXAMPLE  60. — Find  the  thickness  of  the  packing  in  a  stuffing  box  for  a  valve-rod, 
the  rod  being  1  inch  in  diameter. 

J"  +  4"  —  i  inch  for  the  thickness  of  the  packing. 

EXAMPLE  61. — Find  the  thickness  of  the  packing  in  a  piston-rod  stuffing  box,  the 
piston-rod  being  3^  inches  in  diameter. 

3i" 

+  i"  =  IfV  inches  for  the  thickness  of  packing. 

In  good  locomotive  practice  we  find  the  average  thickness  of  packing  for  rods 
of  small  diameter  £  inch,  and  for  large  piston-rods  1&  inches.  Here  we  see  that 
the  dimensions  obtained  by  the  rule  agree  with  practice.  If  the  rods  are  larger  than 
3J  inches  in  diameter,  this  rule  cannot  be  used,  because  the  thickness  of  the  packing 
so  found  will  be  too  great.  But,  since  locomotive  piston-rods  are  seldom  larger  than 
3£  inches  in  diameter,  we  may  conclude  that  the  foregoing  rule  can  be  used  for  finding 
the  thickness  of  the  packing  in  all  stuffing  boxes  iised  in  locomotives. 

As  soon  as  we  know  the  thickness  of  the  packing,  the  diameter  7  of  the  stuffing 
box  is  readily  obtained,  for  we  have  only  to  add  twice  the  thickness  of  the  packing  to 
the  diameter  of  the  rod ;  the  sum  will  be  the  diameter  /. 

194.  Generally  speaking,  the  greater  the  depth  H  of  the  box  the  longer  the 
engine  will  run  without  renewing  the  packing.     In  locomotives  the  depth  //  of  the 
piston-rod  stuffing  boxes  is  often  limited,  therefore  the  general  practice  is  to  make  this 
depth  equal  to  1^  to  1J  times  the  diameter  I  of  the  stuffing  box ;  this  proportion  of 
diameter  to  the  depth  of  the  box  is  also  adopted  for  the  valve-rod  stuffing  boxes. 

195.  Stuffing-box  glands  for  the  valve-rod  are  sometimes  made  of  brass,  and  many 
are  made  of  cast-iron ;  piston-rod  glands  are  nearly  always  made  of  cast-iron. 

In  order  to  reduce  the  number  of  patterns  as  much  as  possible,  and  also  to  reduce 
the  number  of  tools  and  templets  necessary  for  boring  and  turning  the  glands,  the 
dimensions  of  a  brass  gland  and  a  cast-iron  gland  are  alike ;  hence  the  following  rules 
apply  to  glands  made  of  either  metal : 

For  valve-rods  of  small  diameter  the  glands  are  often  made  entirely  of  brass,  as 
shown  in  Fig.  266.  When  the  larger  glands  are  made  of  cast-iron,  they  are  lined  with 
a  brass  bushing  (Fig.  265)  in  the  same  way  as  all  large  glands  are  lined  where  the  cost 


MODERN  LOCOXOTirE  COXSTRrCTIOX. 


179 


of  labor  in  making  the  bushing  is  less  than  the  cost  of  making  the  gland  entirely  of 
brass.  In  the  smaller  oast-iron  glands  the  bushing  is  generally  £  inch  thick,  and 
in  larger  glands  ^  inch  thick.  The  end  b.2  of  the  bushing  is  enlarged;  some- 
times this  enlarged  part  will  cover  the  whole  end  of  the  gland  as  shown,  and  this  the 
writer  believes  to  be  the  best  practice;  at  other  times  the  diameter  of  this  end  is 
only  a  very  little  larger  than  the  diameter  of  the  bushing,  the  cast-iron  part  being 
counterbored  to  receive  the  enlarged  end.  The  bushing  is  forced  into  the  cast-iron ; 
its  object  is  to  prevent  the  collection  of  rust  on  the  inside  of  the  gland  when  the 
engine  stands  still  for  any  considerable  length  of  time,  and  thus  prevent  scratching  or 
injuring  the  rod. 

196.  For  finding  the  thickness  g  of  the  flange  on  gland,  Fig.  265,  we  may  use  the 
following  rule : 

RULE  34. — To  J  of  the  diameter  of  the  rod  add  f  of  an  inch ;  the  sum  will  be  the 
thickness  g  of  the  flange. 

EXAMPLE  62. — Find  the  thickness  of  the  flange  on  a  piston-rod  gland,  the  rod 
being  3  inches  in  diameter. 

|"  +  |"  =  i£  inches  for  the  thickness  of  the  flange. 

The  flanges  on  the  piston-rod  gland  are  generally  made  oblong  in  form,  as  shown 
in  Fig.  268 ;  sometimes  oil  chambers  are  cast  in  these  flanges,  as  shown  in  Fig.  269 ; 
the  thickness  of  the  latter  flanges  is  somewhat  greater  than  the  thickness  obtained  by 


Fly.  27O 


Fig.  268 


Fly.  269 


the  rule.  The  flanges  on  the  valve-rod  glands  are  generally  circular  in  form,  as  shown 
in  Fig.  270.  The  length  (Fig.  268)  of  the  piston-rod  gland,  and  the  diameter  of  the 
valve-rod  gland  (Fig.  270)  must  be  made  sufficiently  great  to  allow  the  edges  of  the 
flanges  to  project  a  little  beyond  the  nuts  of  the  studs. 

The  length  of  the  gland  measured  from  the  flange  to  the  end  is  generally  made 
equal  to  if  or  £  of  the  depth  H  of  the  stuffing  box. 

197.  A  brass  ring  d,  as  shown  in  Fig.  l2<>.">,  is  placed  inside  of  the  stuffing  box. 
The  hole  in  this  ring  is  just  large  enough  to  allow  the  piston-rod  to  pass  through 
easily,  whereas  the  hole  e  through  the  cylinder  head,  or,  in  other  words,  the  hole  e 
through  that  portion  of  the  cast-iron  which  forms  one  end  of  the  stuffing  box,  is 
generally  made  J  inch  larger  in  diameter  than  that  of  the  piston-rod.  This  ar- 
rangement prevents  iron  and  iron  from  touching  each  other,  and  consequently  what 


180  MODERN  LOCOMOTIVE   CONSTRUCTION. 

little  rust  may  form  and  collect  in  tlie  hole  e  through  the  cylinder  head  cannot  scratch 
or  injure  the  rod.  Sometimes  we  find  the  brass  ring  d  extending  through  the 
cylinder  head ;  in  cases  of  this  kind  the  form  of  the  ring  d  will  be  similar  to  that  of 
the  ring  d  shown  in  Fig.  266.  Allowing  the  ring  d  to  pass  through  the  cylinder  head 
in  this  manner  is,  in  the  writer's  opinion,  bad  practice,  because  this  ring  is  liable  to 
break,  and  sometimes  it  does  break;  then  if  pieces  of  it  fall  into  the  cylinder, 
more  or  less  damage  to  the  cylinder  head  or  piston  may  be  the  result ;  the  writer  has 
known  accidents  of  this  kind  to  happen,  incurring  a  great  expense  for  repairing  the 
damage.  Although  brass  rings  d  similar  in  form  to  that  shown  in  Fig.  266  should 
not  be  used  in  any  kind  of  stuffing  boxes,  the  use  of  this  form  of  ring  cannot  be 
avoided  in  a  valve-rod  stuffing  box  for  the  following  reason : 

The  valve-stem  is  forged  to  the  valve  yoke,  and  therefore  it  must  be  entered  into 
the  stuffing  box  from  the  inside  of  the  steam  chest;  this  cannot  be  done  without 
canting  the  valve-stem,  because  the  form  of  the  steam  chest  will  not  allow  it  to  enter 
the  hole  squarely,  and  consequently  this  hole  must  be  considerably  larger  in  diameter 
than  that  of  the  valve-stem;  when  the  stem  is  in  position  the  hole  is  reduced  by 
the  brass  ring  d,  hence  this  shape.  From  the  fact  that  a  large  hole  is  required  for 
entering  the  valve-stem  into  the  stuffing  box,  it  will  be  readily  perceived  that  the 
valve-stem  cannot  be  taken  out  of  the  steam  chest  or  placed  into  the  same  without 
first  removing  the  ring  d,  and  therefore  this  ring  must  not  be  fitted  very  tightly  in  the 
stuffing  box. 

The  face  of  the  i-ing  d  and  the  end  of  the  gland  which  touches  the  packing  are 
generally  turned  slightly  concave,  so  as  to  help  to  force  the  packing  against  the  rod. 

198.  To  keep  the  packing  in  place,  and  to  compress  it  sufficiently  to  prevent 
leakage,  the  gland  must  be  forced  against  the  packing,  and  for  this  purpose  the  studs 
c  c  (Figs.  265  and  266)  are  used.  In  the  piston-rod  stuffing  box  two  studs  are 
generally  employed;  the  limited  space  for  the  gland  prevents  the  use  of  a  greater 
number  of  studs.  In  the  valve-rod  stuffing  box  two  or  three  studs  are  used,  the 
number  of  studs  depending  on  the  fancy  and  judgment  of  the  designer. 

In  the  piston-rod  stuffing  box  the  studs  should  be  placed  sufficiently  far  apart  to 
allow  the  hub  of  the  crosshead  to  pass  between  the  nuts  on  these  studs.  In  the  valve- 
rod  stuffing  box  the  distance  from  the  center  of  the  studs  to  the  center  of  the  box 
should  be  sufficiently  great  to  allow  the  tap  to  pass  the  outside  of  the  box  when  the 
holes  for  these  studs  are  being  tapped. 

For  finding  the  diameter  of  these  studs. when  two  are  used  in  each  box  we  have 
the  following  rule : 

RULE  35. — To  J  of  the  diameter  of  the  rod  add  ^  inch;  the  sum  will  be  the 
diameter  of  the  stud.  Putting  this  rule  in  the  shape  of  a  formula,  we  have, 

4  diameter  of  rod  +  4  inch  =  diameter  of  stud. 

EXAMPLE  63. — Find  the  diameter  of  the  studs  for  a  piston-rod  stuffing  box,  the 
rod  being  3  inches  in  diameter. 

f "  +  J"  =  1  inch  for  the  diameter  of  the  stud. 

The  same  rule  may  be  used  in  finding  the  diameter  of  the  studs  for  the  valve-rod 
stuffing  box  when  three  studs  are  used  for  each  box ;  although,  theoretically,  when 


MODERN  LOCOMOTirE   COXSTRUCTIOX. 


181 


more  than  two  studs  are  employed,  the  diameter  of  each  stud  can  be  made  somewhat 
less  than  that  found  by  the  rule. 

The  general  practice  is  to  make  all  studs  for  small  piston-rod  stuffing  boxes  £ 
inch  in  diameter;  for  piston-rod  stuffing  boxes  on  cylinders  12  inches  in  diameter 
and  up  to  17  inches  in  diameter  £"  studs  are  used;  for  cylinders  18  to  20  inches 
in  diameter  1"  studs  are  used;  and  for  larger 
cylinders  Ij"  studs.  For  the  valve-rod  stuffing 
boxes  the  diameter  of  the  studs  varies  from 
f  to  J  inch,  according  to  the  size  of  the  rod. 

Two  nuts  are  always  used  for  each  stud; 
often  both  nuts  are  placed  outside  of  the  gland, 
as  shown  in  the  illustrations,  but  sometimes 
the  nuts  are  placed  so  that  one  will  be  outside, 
and  the  other  nut  inside  of  the  flange  g. 

These  nuts  are  case  hardened  to  prevent 
their  corners  from  becoming  worn ;  and  for  the 
sake  of  convenience  (so  that  one  wrench  can 
be  used)  the  nuts  for  the  piston-rod  gland  and 
those  for  the  valve-rod  gland  are  made  the  same  size,  even  when  larger  studs  are  used 
for  the  former  than  for  the  latter. 

When  only  two  studs  are  employed  great  care  must  be  taken  to  have  the  center 
of  the  studs  and  the  center  of  the  rod  in  one  straight  line,  otherwise  trouble  will  be 
experienced  in  screwing  up  the  gland ;  in  fact,  when  the  two  studs  cannot  be  placed 
exactly  in  line  with  the  center  of  the  rod  it  is  better  to  use  three  studs. 

Another  manner  of  compressing  the  packing  is  shown  in  Fig.  267.  Instead  of 
having  a  gland,  we  have  here  a  brass  sleeve  a  fitting  the  rod  and  the  inside  of  the 
stuffing  box.  This  sleeve  is  pressed  against  the  packing  by  means  of  the  nut  6, 
which  is  tapped  to  fit  the  thread  cut  on  the  outside  of  the  stuffing  box  d.  It  will  be 
noticed  that  the  sleeve  a  has  a  small  flange  on  top;  the  purpose  of  this  flange  is 
simply  to  prevent  the  sleeve  from  being  pressed  too  far  into  the  box,  and  to  provide 
some  means  of  pulling  it  out  of  the  box. 


CHAPTER    V. 

FRAMES.— AXLE  BOXES. 
FRAME  PEDESTALS. 

199.  The  figures  numbered  2*71  up  to  279  represent  the  proportions  of  frame 
pedestals  for  passenger  locomotives  of  different  sizes. 

The  function  of  the  pedestal  is  to  hold  the  axle  box — often  called  the  driving 
box — at  a  given  distance  from  the  cylinder,  but  in  the  meantime  allowing  the  axle 
box  to  move  in  a  vertical  direction.  The  pedestals  are  made  of  wrought-iron,  and 
each  one,  as  shown  in  Figs.  271  and  280,  consists  of  a  portion  of  the  upper  frame 
brace  B,  the  pedestal  legs  A  A2,  and  the  mechanism  used  for  preventing  an  increase 
of  the  opening  at  the  bottom  of  the  jaw.  The  three  parts,  namely,  the  portion  B  of 
the  frame  brace  and  the  two  pedestal  legs  A  A.z,  form  what  is  called  the  pedestal  jaw. 
The  pedestal  legs  in  large  locomotives  are  often  connected  at  the  bottom  by  the 
bolt  D  passing  through  a  frame  thimble  T  inserted  in  the  opening  of  the  jaw,  as 
shown  in  Fig.  271 ;  in  smaller  engines  the  legs  are  connected  by  a  pedestal  cap  C  as 
shown  in  Figs.  277,  278,  279,  and  also  in  280.  The  frame  thimbles  T  are  made  of  cast- 
iron,  the  caps  C  are  made  of  wrought-iron.  The  reason  for  not  using  the  bolts  D 
and  frame  thimbles  in  smaller  locomotives  is  that  the  bolt  D  will  interfere  with  the 
wedge  bolt  E,  as  will  be  presently  explained. 

The  pedestals  are  united  in  each  frame  by  the  upper  frame  brace  B2  and  the 
lower  frame  brace  L ;  these  braces  are  forged  to  the  pedestal  jaws. 

There  are  two  distinct  forms  of  pedestal  jaws  used ;  one  is  represented  in  Fig.  271 
and  the  other  in  Fig.  280.  The  difference  between  these  two  forms  consists  in  the  shape 
of  one  of  the  pedestal  legs,  thus :  The  jaw  represented  in  Fig.  271  has  one  straight 
and  one  tapered  leg — that  is  to  say,  the  inside  of  one  of  the  pedestal  legs,  as  A2  (called 
the  straight  leg),  is  planed  square  with  the  top  of  frame,  the  inside  of  the  other 
one,  A,  is  planed  so  as  to  form  an  angle  with  the  top  of  frame,  making  the 
opening  at  the  bottom  of  the  jaw  greater  than  at  the  top.  In  the  pedestal  rep- 
resented in  Fig.  280  both  legs  are  tapered.  The  form  of  pedestal  shown  in  Fig. 
280  has  been  used  very  extensively  in  former  years;  but  lately  the  use  of  the  form 
shown  in  Fig.  271  has  increased.  Some  master-mechanics  prefer  to  use  for  all 
locomotives  the  pedestal  caps  as  shown  in  Figs.  277,  278,  and  279  in  place  of  the 
cast-iron  frame  thimbles  T  and  the  bolts  D;  but  the  writer  believes  that  the  use 
of  the  thimble  and  bolt  will  add  to  the  stiffness  of  the  pedestal,  because  by  the  use 
of  the  bolt  D  the  lower  frame  brace  L  will  be  brought  nearer  in  line  with  the 


LOCOMOTIVE  CONSTRUCTION. 


183 


r-f-tt n  i 


...  * Lf 


184 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


>*»• 


center  of  driving  axle;  by  so  doing  a  better  dis- 
tribution of  metal  in  the  frame  is  secured  and  the 
stiffness  of  the  pedestal  legs  increased. 

WEDGES. 

200.  The   function   of  the  wedges — shown  in 
section,  and  marked  W  W2  in  Fig  271 — is  twofold: 
first,  they  protect  the  pedestal  legs  from  wear ;  sec- 
ondly, with  these  wedges  the  play  is  taken  up  be- 
tween the  axle  box  and  the  wedges  caused  by  the 
wear,  which  will  result  from  the  vertical  movement 
of  the  axle  box.     The  wedge  marked  W  is  called 
the  "  short  wedge,"  and  that  marked  W2  is  called 
the  "long  wedge."    It  may  here  be  necessary  to 
remark  that  the  long  wedge  W2  in  this  pedestal  has 
in  nowise  a  wedge   shape,  consequently  "a  shoe" 
would  be  a  better  name  for  it;  we  shall  for  the 
sake  of  simplicity  follow  the  usual  custom  and  re- 
tain  the  term  "wedge,"  as  this  term  will  cover 
both  wedges  used  in  the  pedestal  shown  in  Fig.  280, 
in  which  the  long  wedge  W2  must  necessarily  have 
the  shape  of  a  wedge  similar  to  that  of  the  short 
one.     The  wedges  must  be  accurately  fitted  in  the 
pedestal  jaw,  so  that  the  wearing  surfaces  s  s  and 
s2  s2  of  these  two  wedges  will  be  exactly  parallel  to 
each  other,  and  perpendicular  to  the  top  of  the 
frame ;  the  distance  between  these  wearing  surfaces 
should  be  equal  to  the  width  of  the  axle  box. 

LONG    WEDGE. 

201.  The  long  wedge  is  always  fitted  to  the 
straight  pedestal  leg,  and  since  the  wearing  surface 
s2  s2  (Fig.  271)  of  the  wedge  must  stand  perpendic- 
ular to  the  top  of  frame,  it  follows  that  the  thick- 
ness of  the  metal  forming  the  wearing  surface  s2  s2 
must  be  the  same  throughout.     The  length  of  the 
wedge  W2  is  equal  to  the  distance  between  the  top 
of  the  thimble  T  and  the  bottom  of  the  frame  brace 
B,  so  that  this  wedge  cannot  move  in   a  vertical 
direction,  and  it  is  further  secured  in  position  by 
the  screw  bolt  R2,  which  holds  the  wedge  firmly 
against  the  straight  leg.     In  order  to  prevent  on 
the  long  wedge  the  formation  of  ridges,  the  thick- 
ness of  the  metal  at  the  ends  is  reduced,  causing 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


185 


the  wearing  surfaces  on  three  sides  of  the  wedge  to  project  beyond  the  metal  at  the 
ends,  and  these  projecting  surfaces  are  accurately  planed.  The  long  wedge  is  shown  in 
•.Mail  in  Figs.  281  and  282. 


SHOUT  WEDGE. 


202.  A  reduction  of  the  metal  at  the  ends  of  the  short  wedge  W  (Fig.  271)  is 
not  necessary,  because  the  length  of  the  wedge  is  always  made  equal  to,  or  a  little  less 
than,  the  length  of  the  axle  box,  and  consequently  ridges  cannot  be  formed  on  it, 
as  the  axle  box  in  its  vertical  movement  will  move  beyond  the  ends  of.  the 


-Bz 


- -• 13V 


Fig.  280 


O 


g.  ssi 


U» 


O 

' 

11 

i 

O 


J'i'j.  SSa       J-'lg.  287 


fig.  XS3  •*''(/•  »* 


Tig.  USB 


Tiff.  HS8 


wedge.  Since  the  wearing  surface  s  s  must  be  perpendicular  to  the  top  of  frame,  and 
since  this  wedge  has  to  fit  the  tapered  pedestal  leg,  it  follows  that  the  part  of  it 
against  which  the  side  of  the  axle  box  slides  must  have  a  wedge  form,  as  shown. 
The  play  between  the  wedges  and  the  axle  box  is  taken  up  by  moving  the  short  wedge 
upwards  by  means  of  the  wedges  bolt  K.  By  so  doing  the  short  wedge  will  slide 
against  the  inner  surface  of  the  tapered  pedestal  leg  A,  and  thereby  reduce  the 
distance  between  the  wedges.  When  the  short  wedge  has  been  adjusted  to  the  correct 
position,  it  is  held  there  by  the  wedge  bolt  E,  and  also  by  the  screw  bolt  R,  which 
holds  it  firmly  against  the  pedestal  leg.  For  the  bolt  R  a  slot  is  cut  in  the  leg,  so 


MODERN  LOCOMOTIVE   CONSTRVCT10X. 

that  the  bolt  can  be  moved  up  or  down  with  the  wedge  to  any  desired  position.    The 
short  wedge  is  shown  in  detail  in  Figs.  283  and  284. 

WEDGE  BOLTS. 

203.  When  a  cast-iron  thimble  is  used,  as  shown  in  Fig.  271,  two  wedge  bolts  Z? 
are  employed  for  each  short  wedge.     These  bolts  pass  through  slots  m  n,  Fig.  273«, 
cast  into  the  thimble  near  the  end,  one  on  each  side  of  the  bolt  D.     The  reason  for 
casting  slots  near  both  ends  of  the  thimble,  as  shown,  is  to  make  the  thimble  revers- 
ible.    The  heads  of  the  wedge  bolts  are  cylindrical  in  form,  as  represented  in  Fig. 
285.     These  heads  fit  into  recesses  //cast  in  the  short  wedge,  as  shown  in  Fig.  283. 
In  small  locomotives  the  frames  are  not  sufficiently  wide  to  admit  two  wedge  bolts 
and  therefore  only  one  can  be  used,  and  this  bolt  must  be   placed  in  the  center 
of  the  wedge,  as  indicated  by  the  recess  /  in  the  wedge  shown  in  Fig.  286.    But 
placing  the  wedge  bolt  in  the  center  of  the  wedge  will  prevent  the  use  of  a  pedestal 
bolt  I),  which  must  also  pass  through  the  center  of  the  pedestal  legs,  and  therefore 
these  two  bolts  will  interfere  with  each  other.     It  is  for  this  reason  that  in  small 
locomotives  wrought-iron  pedestal  caps  (7,  as  shown  in  Fig.  277,  are  employed  in 
place  of  the  cast-iron  thimble  T  and  pedestal  bolts  D. 

In  the  writer's  opinion  it  is  always  best  not  to  tap  the  pedestal  caps  for  the  wedge 
bolts  E,  but  to  allow  this  bolt  to  pass  through  a  slot  cut  in  the  pedestal  cap.  In 
this  case  the  wedge  bolt  will  have  the  same  form  of  head  as  shown  in  Fig.  285.  If  the 
pedestal  cap  is  tapped  for  the  wedge  bolt,  then  the  head  of  this  bolt  must  have  a 
conical  form,  as  shown  in  Fig.  288. 

PROPOKTIONS  OF  WEDGES  AND  BOLTS. 

204.  We  have  already  stated  that  the  length  of  the  long  wedge  must  be  equal  to 
the  length  of  opening  in  the  pedestal ;  and  the  length  of  the  short  wedge  equal  to  the 
length  of  the  axle  box,  or  a  little  shorter.     In  the  long  wedge  the  thickness  of  the 
metal  which  forms  the  wearing  surface  s2  s^  should  not  be  less  than  that  given  in  the 
illustrations ;  and  the  thinnest  part  of  the  short  wedges  should  be  about  £  inch  for 
small  locomotives,  and  f  inch  for  large  ones.     The  flanges  of  all  the  wedges  in  small 
locomotives  should  not  be  less  than  f  inch  thick,  and  for  larger  ones  l£  inches  thick ; 
the  exact  thickness  of  these  flanges  for  the  different  sizes  of  axle  boxes,  and  which 
the  writer  would  recommend,  are  given  in  Figs.  316  to  340.     The  diameter  of  the 
wedge  bolts  E  is  usually  £  inch  for  the  small  locomotives,  and  I  inch  for  larger  ones. 
The  bolts  which  hold  the  wedges  to  the  pedestal  legs  are  generally  made  £  inch  in 
diameter. 

205.  All  the  principal  dimensions  of  the  pedestals  for  passenger  locomotives  are 
given  in  our  illustration,  and  these  dimensions  agi'ee  with  modern  locomotive  practice. 
In  connection  with  this  subject  it  may  here  be  remarked  that  in  late  years  the 
mechanisms  of  the  larger  locomotives  have  been  made  heavier  and  their  weights 
increased.     It  will   therefore  be  found  by  comparison   that  the  dimensions  of  the 
larger  pedestals  will  exceed  the  dimensions   of  pedestals  made  a  few  years  ago, 
and  the  dimensions  given  for  the  smaller  pedestals  will  agree  very  closely  with  the 


MOI)K1!\  LOCOMOTIVE   CONSTRUCTION. 


dimensions  of  the  average  pedestals 
now  in  use.  The  dotted  circle,  with 
the  diameter  given  in  each  ped- 
estal, represents  the  driving  axle 
journal  suitable  for  each  one  of 
these  pedestals.  It  will  be  noticed 
thtit  the  diameter  of  the  journal 
given  in  Fig.  271  is  considerably 
larger  than  the  average  diameter  of 
journals  used  in  passenger  locomo- 
tives built  some  years  ago ;  but  in 
modern  engines  of  this  class  jour- 
nals as  large  as  shown  in  this  fig- 
ure are  now  used,  and  the  writer 
believes  it  is  only  a  matter  of  time 
when  this  size  of  axle  will  be  gen- 
erally adopted  for  fast  passenger 
engines  having  cylinders  18  inches 
in  diameter. 

ENGINE  FRAMES. 

206.  Fig.  289  represents  the 
main  frame  for  an  eight-wheeled 
passenger  engine,  such  as  is  shown 
in  Fig.  1,  and  suitable  for  a  locomo- 
tive having  cylinders  18  inches  in 
diameter.  Fig.  290  represents  the 
front  splice  of  the  same  frame ;  the 
front  splice  is  fastened  to  the  main 
frame,  as  shown  in  Fig.  289,  in 
which  that  portion  marked  S  rep- 
resents one  end  of  the  front  splice. 

Fig.  291  represents  the  main 
frame  for  a  locomotive  of  the  same 
class  as  the  foregoing,  but  having 
cylinders  10  inches  in  diameter. 

The  back  ends  of  these  main 
frames,  Figs.  289  and  291,  aiv  suit- 
able for  a  footboard,  and,  since 
nearly  all  locomotives  which  carry 
footboards  burn  soft  coal  or  wood, 
it  may  be  said  that  these  frames 
are  for  soft  coal  and  wood  burning 
locomotives.  For  this  class  of  lo- 
comotives the  horizontal  distance 


OS 

8 
* 


i 

I 


188  MODERN  LOCOMOTIVE  CONSTRUCTION. 

from  the  center  of  the  rear  axle  to  the  back  end  of  frame  is  usually  42  inches.  For 
hard  coal  burning  locomotives  this  distance  may  have  to  be  changed,  and  made  either 
longer  or  shorter  to  suit  the  design  of  boiler. 

In  designing  a  locomotive  frame  the  first  step  is  to  locate  the  centers  of  the 
driving  wheels  and  the  position  of  the  cylinders.  It  may  be  said  that  the  relative 
position  of  the  driving  wheels  and  the  cylinder  depend  upon  the  proper  distribution 
of  the  weight  on  the  drivers,  and  also  on  the  length  of  the  boiler.  Again,  in  all  eight- 
wheeled  passenger  engines,  such  as  shown  in  Fig.  1,  ten-wheeled  engines,  shown  in 
Fig.  2,  and  mogul  engines,  shown  in  Fig.  3,  which  are  designed  for  burning  soft  coal 
or  wood,  the  fire-box  is  placed  between  the  two  rear  axles,  and  consequently  the 
distance  between  these  axles  must  be  sufficiently  great  to  admit  the  fire-box  between 
them ;  there  must  also  be  sufficient  room  for  the  working  of  the  eccentrics,  space  for 
the  axle  boxes,  room  enough  for  cleaning  the  water  space  around  the  furnace,  and 
such  space  as  may  be  required  for  other  special  mechanism  which  the  design  of  the 
locomotive  may  call  for.  But  in  the  meantime  it  must  be  remembered  that  the  distance 
between  the  centers  of  any  two  wheels  which  are  connected  by  a  side  rod  must  not 
exceed  8'  9"  or  9'  0"  at  the  utmost ;  the  latter  distance  is  seldom  used.  If  the  distance 
between  the  centers  of  the  driving  wheels  exceed  these  distances,  the  length  of  the  side 
rods  will  become  too  great,  and  consequently  dangerous ;  because,  on  account  of  the  great 
number  of  revolutions  per  minute  of  the  driving  wheels,  the  change  of  motion  of  the 
side  rods  from  an  upward  to  a  downward  or  from  a  downward  to  an  upward  motion 
becomes  so  sudden  that  the  weight  of  the  rods  will  be  an  element  of  danger,  causing 
the  side  rods  which  are  longer  than  8'  9"  or  9'  0"  to  Tae  shaken  to  pieces.  From  these 
remarks  we  learn  that  in  the  classes  of  locomotives  before  mentioned  the  greatest 
distance  between  the  center  of  the  rear  driving  wheel  and  the  center  of  the  one  next  to 
it  is  limited  by  the  length  of  the  side  rod,  and  the  shortest  distance  between  the  centers 
of  the  same  drivers  in  the  same  classes  of  engines  is  limited  by  the  length  of  the  fire- 
box. 

In  ten-wheeled  locomotives  the  distance  between  the  center  of  the  middle  driving 
wheel  and  the  center  of  the  front  one  depends  greatly  upon  the  general  design  of 
the  engines ;  but  usually  the  position  of  the  front  drivers  in  these  engines  is 
determined  by  that  of  the  front  truck,  and  sometimes  by  the  valve  motion.  In  all 
ten-wheeled  engines  that  have  come  under  the  writer's  notice,  the  distance  between 
the  centers  of  the  middle  and  front  drivers  has  been  less  than  that  between  the  centers 
of  the  rear  and  middle  drivers. 

In  mogul  engines,  the  front  driving  wheels  are  generally  placed  as  far  forward  as 
the  cylinder  will  permit,  leaving  just  room  enough  for  removing  the  cylinder  head  and 
casing  without  striking  the  tire.  In  these  engines,  too,  the  distance  between  the 
centers  of  the  middle  and  front  drivers  is  generally  less  than  that  between  the  centers 
of  the  middle  and  rear  drivers. 

In  consolidation  we  have  the  first,  second,  third,  and  fourth  pair  of  driving 
wheels ;  the  pair  of  driving  wheels  next  to  the  cylinder  is  called  the  first  pair.  The 
same  conditions  which  determine  the  position  of  the  front  drivers  in  a  mogul  engine 
will  also  determine  the  position  of  the  first  pair  of  drivers  in  the  consolidation  engine ; 
that  is,  in  these  engines  the  front  drivers  are  placed  as  far  forward  as  the  cylinders 


MODERX  LOCOMOTIVE  CONSTRUCTION. 

will  permit,  so  that  the  cylinder  head  and  casing  can  readily  be  taken  off.  The 
distance  between  the  centers  of  the  first  and  second  pair  of  drivers  must  be  sufficiently 
great  to  admit  the  rocker  between  the  tires  of  these  wheels.  The  distances  between 
the  centers  of  the  second  and  third  pair,  and  between  the  third  and  the  fourth 
pair,  are  generally  arranged  so  as  to  leave  1  inch  or  1J  inches  clearance  between  the 
nanges  of  the  tires. 

207.  In  small  locomotives  the  total  wheel  base  generally  depends  on  the  proper 
distribution  of  the  weight  of  the  engine  on  all  the  wheels.    For  instance,  moving  the 
front  truck  nearer  to  or  further  from  the  center  of  gravity  of  the  locomotive,  we 
throw  more  or  less  weight  on  the  truck.    In  larger  engines  we  may  often,  if  it  is 
desirable,  be  able  to  move  the  front  truck  nearer  to  the  center  of  gravity  of  the  loco- 
motive ;  but  if  we  attempt  to  move  the  truck  away  from  the  center  of  gravity  of  the 
engine,  we  may  meet  with  obstacles,  namely,  the  sharp  curves  of  the  track  over  which 
the  engine  has  to  run,  and  for  which  the  wheel  base  must  be  kept  as  short  as  possible. 
The  turn-tables  of  the  road  may  also  limit  the  length  of  the  wheel  base.     Therefore  it 
will  be  seen  that  the  arrangement  of  the  wheels,  and  the  determination  of  the  total 
wheel  base  of  large  locomotives,  is  brought  within  very  narrow  limits.     And  it  may 
be  said  that,  in  cases  of  this  kind,  the  ingenuity  of  the  designer  is  often  taxed  to  the 
utmost  to  obtain  satisfactory  results ;  and  even  then  he  may  have  to  be  satisfied  with 
results  not  as  desirable  as  they  should  be. 

208.  The  relative  positions  of  the  wheels  under  hard-coal  burners  are  some- 
times the  same  as  those  under  soft-coal  burners ;  at  other  times  conditions  will  arise 
which  will  compel  a  change  in  the  arrangement  of  the  wheels  under  the  hard-coal 
burners. 

DEPTH   OF  PEDESTAL. 

209.  The  depth  of  the  pedestal— that  is,  the  distance  D,  Fig.  289,  from  the  top  of 
the  cast-iron  thimble  to  the  under  side  of  the  upper  frame  brace  B — should  be  suf- 
ficient to  allow  the  driving  box  to  move  a  given  amount  in  a  vertical  direction,  thus: 
In  Fig.  289  the  line  marked  F  represents  the  top,  and  the  line  marked  G,  the  bottom 
of  driving  box.     The  depth  of  the  space  between  the  top  of  the  box  and  the  frame 
brace  B,  plus  the  depth  of  the  space  between  the  bottom  of  the  box  and  the  thimble, 
represents  the   total  vertical   movement  of  the  driving  box.     When   a  locomotive 
is  in  good  working  order,  with  the  usual  amount  of  fuel  and  water,  the  driving  box 
should  occupy  in  the  pedestal  a  position  in  which  the  upper  clearance — that  is,  the 
space  between  the  top  of  box  and  frame  brace — is  greater  than  the  lower  clearance, 
or  the  space  between  the  bottom  of  the  box  and  thimble.     Thus:  In  Fig.  289  we 
see  that  the  upper  clearance  is  3  inches,  and   lower   clearance  is  1A   inches.     The 
total  amount  of  clearance  and  the  difference  between   the  top  and  bottom  clear- 
ance is  arbitrary  and  is  not  always  alike  in  the  same  class  of  locomotives.     The 
average  amount  of  clearance  at  the  top  and  bottom  of  the  boxes  for  the  different 
sizes   of    locomotives,   as  generally   adopted    by  locomotive   builders    and   master- 
mechanics,  is  given  in  Figs.  271  to  279,  in  which  the  dotted  lines  immediately  over 
and  under  the  axle  represent  the  top  and  bottom  of  the  driving  boxes;  the  dimen- 
sion given  from   the  top  of  the  box   to  frame  brace  B  represents  the  amount  of 


190  MODERN  LOCOMOTIVE    CONSTRUCTION. 

the  upper  clearance,  and  the  dimension  given  from  the  bottom  of  the  box  to  the 
thimble  or  pedestal  cap  represents  the  amount  of  the  lower  clearance. 


WIDTH  OF  PEDESTAL  OPENING. 

210.  The  width  of  the  opening  of  the  pedestal,  or  the  distance  from  leg  to  leg, 
Fig.  289,  should  be  such  as  will  not  admit  the  short  wedge   further  into  the  ped- 
estal after  the  driving  box  and  long  wedge  are  in  position  than  is  necessary  for  it 
to  clear  the  wedge-bolt  nut  on  the  top  of  the  thimble,  leaving  as  great  a  distance  as 
possible  between  the  top  of  the  short  wedge  and  the  frame  brace  B,  through  which  the 
short  wedge  can  be  moved  to  take  up  the  play.     The  distance  H  given  in  Fig.  289, 
from  the  center  line  x  Y  to  the  face  of  the  short  wedge,  represents  one-half  the  width 
of  the  driving  box ;  and  so  also  the  dimensions  from  the  vertical  center  lines  to  the 
face  of  the  short  wedges  in   Figs.  271   to  279  represent  one-half  the  width  of  the 
axle  boxes. 

TAPEK   OF  PEDESTAL  LEGS,   AND  POSITION  OF   STKAIGHT  LEG. 

211.  When  pedestals  are  used  like  those  shown  in  Figs.  289  and  291,  the  straight 
leg  should  always  be  placed  towards  the  cylinder;  by  so  doing  the  distance  from  the 
cylinders  to  the  center  of  the  driving  wheels  cannot  be  readily  changed,  and  therefore 
the  distance  from  center  to  center  of  the  brasses  in  the  main  rod  need  not  be  so  often 
adjusted.     The  amount  of  taper  for  the  inner  surface  of  the  tapered  legs  is  generally 
l£  inches  in  12  inches ;  and  this  taper  is  used  for  all  pedestals  of  the  form  shown  in 
Figs.  289,  291,  and  280. 

POSITION   OF  CENTEK  LINES. 

212.  In  connection  with  this  subject  it  may  be  advantageous  to  the  reader  to  call 
his  attention  to  the  fact  that  when  pedestals  such  as  shown  in  Figs.  289  and  291  are 
used,  the  vertical  center  line  drawn  through  the  center  of  the  axle  does  not  pass 
through  the  center  of  the  opening  of  pedestal  at  the  top ;  that  is  to  say,  the  distance 
K  (Fig.  289)  from  the  center  line  x  Y  to  the   straight  leg  will  be  greater  than  the 
distance  /  from  the   center  line  a;  I7"  to  the  top  of  the  tapered  leg ;  at   the   bottom 
of  the  pedestal  the  conditions  are  reversed.    When  pedestals  such  as  shown  in  Fig. 
280  are  used,  the  vertical  center  line  drawn  through  the  center  of  the  axle  will  pass 
through  the  center  of  the  opening  of  pedestal,  both  at  the  top  and  bottom. 

It  is  well  to  note  these  facts,  because  in  designing  a  frame  the  position  of 
these  center  lines  have  a  very  important  bearing  in  determining  the  position  and 
dimensions  of  other  parts  of  the  locomotive.  Hence  in  designing  a  frame  having 
pedestals  as  shown  in  Figs.  289  and  291,  the  distance  from  the  straight  pedestal  leg 
to  the  vertical  center  line  x  Y  must  be  equal  to  the  thickness  of  the  long  wedge  added 
to  one-half  the  width  of  the  driving  box,  whereas  for  the  pedestals  shown  in  Fig.  280, 
the  vertical  center  line  must  be  drawn  through  the  center  of  the  opening  of  the 
pedestal.  In  all  pedestals  the  horizontal  center  line  drawn  through  the  center  of  the 
axle  must  be  in  a  position  which  will  give  about  the  same  relative  clearance  on  top 


MODERN  LOCOMOTIVE  CONSTRUCTION.  191 

and  bottom  of  driving  box  as  given  in  Figs.  271  to  279 ;  or,  in  other  words,  in  design- 
ing a  locomotive  frame  the  driving  boxes  must  be  drawn  in  the  same  positions  as 
they  would  occupy  when  the  engine  is  in  first-class  working  order  and  running  on 
the  road;  and  the  boxes  must  be  considered  to  be  stationary  during  the  time  the 
locomotive  is  being  designed. 

WIDTH  OF  FRAME. 

213.  For  small  locomotives  the  width  of  frame  should  not  be  less  than  3  inches, 
so  as  to  provide  on  top  of  frame  a  surface  sufficiently  wide  to  which  the  rocker  box, 
lifting-shaft  bearings,  and  other  mechanism  can  be  bolted  without  interfering  with  the 
necessary  strength  of  the  frame.    In  large  locomotives  the  space  between  the  driving 
wheels  and  the  fire-box  will  limit  the  width  of  the  frame,  and  is  seldom,  if  ever,  wider 
than  4  inches.     We  may  therefore  conclude  that  the  width  of  locomotive  frames 
ranges  from  3  to  4  inches ;  the  suitable  width  of  frame,  such  as  is  usually  adopted 
for  any  one  of  the  different  sizes  of  passenger  locomotives,  will  be  found  in  Figs. 
271  to  279. 

.x1 

DIMENSIONS  OF   FRAME  BRACES. 

214.  To  the  upper  frame  brace  B.,  in  Fig.  292  (also  see  Figs.  271  and  274)  are 
bolted  and  attached  some  of  the  principal  parts  of  the  locomotive.    The  forces  acting 
upon  the  braces  are  of  a  complex  character,  and  therefore  to  find  the  exact  dimensions 
of  the  braces,  which  will  give  them  the  required  strength — no  more  and  no  less — to 
resist  the  forces  acting  upon  them,  would  be  a  very  difficult  matter. 

Consequently,  rules  which  are  to  be  of  any  practical  value  for  finding  the  dimen- 
sions of  a  locomotive  frame,  can  only  be  empirical  or  arbitrary  rules. 

The  following  rules  are  founded  upon  the  observation  of  the  writer,  and  he 
believes  that  the  results  obtained  by  them  will  agree  with  the  best  practice. 

We  have  already  established  the  width  of  the  frames  for  the  various  sizes  and 
classes  of  locomotives ;  our  next  step  will  be  to  find  the  cross-sectional  area  of  the 
upper f runic  brace  !'>.,. 

One  of  the  principal  forces  to  which  locomotive  frames  are  subjected  is  the  pulling 
force,  or  the  horizontal  force,  which  acts  parallel  to  the  frame  braces  B.2  and  L.  This 
pulling  force  is  not  equal  to  the  total  steam  pressure  on  the  piston,  but  for  the  sake 
of  simplicity  in  establishing  the  following  rules,  and  for  convenience  in  finding  the 
dimensions  of  other  locomotive  frames,  we  may  assume  it  to  be  so,  without  falling 
into  any  serious  errors.  Therefore  we  will  again  assume,  as  before,  that  the  maximum 
steam  pressure  in  the  cylinder  is  120  pounds  per  square  inch.  Comparing  the  total 
steam  pressure  on  the  piston  with  the  cross-sectional  area  of  the  frame  brace  I32  in  the 
frames  lately  made,  we  find  that  when  the  cylinders  are  11  inches  and  up  to  18  indies 
in  diameter,  then  1  square  inch  for  every  2,000  pounds  of  the  total  steam  pressure  on 
the  piston  is  allowed  in  the  cross-sectional  area  of  the  frame  braco  /?.,;  for  cylinders 
19  and  20  inches  in  diameter,  1  square  inch  for  every  2,200  pounds;  for  cylinders  •_'] 
and  22  inches  in  diameter,  1  square  inch  for  every  2,400  pounds;  and  for  cylinders  Hi 


192 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


-IB- 
S' 


inches  and  less  in  diameter,  1  square  inch  for  every  1,700  pounds 
of  the  total  steam  pressure  on  the  piston  is  allowed. 

According  to  these  figures,  it  will  be  seen  that  the  cross-sec- 
tional area  of  the  smaller  braces  is  greater  than  that  of  the  larger 
braces,  when  these  are  compared  with  their  respective  piston  press- 
ures. This  is  as  it  should  be,  because  the  ratio  between  the  depth 
and  width  of  the  smaller  braces,  or,  we  may  say,  the  distribution 
of  the  metal  in  the  smaller  braces,  is  not  as  good  for  obtaining 
the  necessary  strength  to  resist  the  forces  which  act  in  a  vertical 
direction  as  the  distribution  of  the  metal  in  or  the  ratio  between 
the  depth  and  breadth  for  the  larger  braces,  and  therefore  the 
larger  frame  braces  can  resist  all  the  forces  acting  upon  them 
with  comparatively  less  metal  than  the  smaller  ones.  It  may  be 
asked :  Why  cannot  we  make  the  form  of  cross-section  in  smaller 
frames  similar  to  that  of  the  larger  ones?  To  this  we  answer 
that  the  widths  of  the  frames  are  determined  by  certain  conditions, 
as  explained  in  Art.  213,  and  cannot  be  changed ;  therefore  if  we 
attempt  to  make  the  depths  of  the  smaller  frame  braces  B2  equal 
or  nearly  equal  to  the  breadth  of  the  frame,  as  is  often  the  case  in 
larger  frames,  we  waste  material  and  obtain  frames  too  heavy  for 
the  smaller  locomotives. 

Hence,  for  finding  the  area  of  those  portions  of  the  upper 
frame  braces  which  are  marked  B.,  in  Tigs.  292,  271,  and  274,  we 
have  the  following  rules : 

EULE  33a. — For  locomotives  having  cylinders  21  or  22  inches 
in  diameter,  divide  the  total  maximum  steam  pressure  in  pounds 
on  the  piston  by  2,400 ;  the  quotient  will  be  the  number  of  square 
inches  in  the  cross-sectional  area  of  the  part  of  the  frame  brace 
marked  B2. 

For  locomotives  having  cylinders  19  or  20  inches  in  diameter, 
divide  the  total  maximum  steam  pressure  on  the  piston  by  2,200. 

For  locomotives  having  cylinders  18  inches  or  less  in  diameter 
down  to  11  inches  in  diameter  (the  latter  included),  divide  the 
total  maximum  steam  pressure  on  the  piston  by  2,000 ;  and  for 
cylinders  10  inches  and  less  in  diameter,  divide  by  1,700;  the 
product  in  each  case  will  be  the  required  area  in  square  inches  of 
the  frame  brace  B2. 

EXAMPLE  64. — What  should  be  the  cross-sectional  area  of  the 
upper  frame  brace  in  a  locomotive  having  cylinders  17  inches  in 
diameter?  Maximum  steam  pressure  on  the  piston  is  120  pounds 
per  square  inch. 

The  area  of  a  piston  17  inches  in  diameter  is  equal  to  226.98 
square  inches ;  hence, 


226.98  x  120 
2000 


=  13.618  square  inches  in  the  sectional  area  of  the  upper  frame  brace. 


MODERN   LOCOMOTIVE   COXSTRrCTIoy.  193 

215.  When  the  area  of  the  upper  frame  brace  B2  is  known  and  the  width  of  the 
frames  established,  as  in  Figs.  271  to  279,  the  depth  of  the  upper  frame  brace  can  be 
readily  obtained,  thus: 

Kn.E  :>4rt. — Divide  the  area  of  the  frame  brace  B2  by  the  suitable  width  of  frame 
given  in  Figs.  271  to  279. 

EXAMPLE  (if). — What  should  be  the  depth  of  the  upper  frame  brace  for  a  locomo- 
tive having  cylinders  18  inches  in  diameter!  Maximum  steam  pressure  on  the  piston 
is  120  pounds  per  square  inch. 

The  area  of  a  piston  18  inches  in  diameter  is  254.47  square  inches ;  hence,  accord- 
ing to  Kule  33«,  the  area  of  the  upper  frame  brace  must  be 

254.47  x  120 

•'(100  :  15.26+  square  inches. 

In  Fig.  271  we  see  that  the  suitable  width  of  the  frame  for  a  locomotive  with 
cylinders  18  inches  in  diameter  should  be  4  inches. 

According  to  Rule  34o,  the  depth  of  the  brace  will  be 

15.2(5 

— r —  =  ,3.81  inches. 

Comparing  this  answer  with  the  dimension  given  in  Fig.  271,  we  find  the  two  to 
agree  very  nearly.  By  the  same  rules  the  depths  B.2  of  all  the  upper  frame  braces  in 
Figs.  272  to  279  have  been  obtained ;  and  in  order  to  avoid  in  these  dimensions  fractious 
of  -fV  inch,  the  depths  of  the  upper  frame  braces  given  in  some  of  these  figures  are 
very  nearly  J  of  an  inch  deeper  than  obtained  by  computation. 

The  part  of  the  upper  frame  brace  marked  B,  which  forms  the  top  of  the  pedestal 
jaw,  is  generally  made  \  inch  deeper — sometimes  more — than  the  depth  of  that  part 
of  the  upper  brace  marked  B.,\  by  so  doing,  the  stiffness  of  the  pedestal  jaw  is 
increased,  and  will  to  some  extent  prevent  injury  to  it  when  the  bolt  D  or  the  pedestal 
cap  C  is  removed. 

The  portion  of  the  upper  frame  brace  marked  _Z?3,  between  the  rear  pedestal  and 
the  rear  end  of  frame,  in  Figs.  289,  291,  292,  is  not  subjected  to  such  severe  vertical 
stress  as  some  of  the  other  portions  of  the  brace,  and  therefore  the  depth  of  that  part 
marked  /?;t  is  generally  made  4  inch  less  than  the  depth  found  by  Rule  ',\4/i. 

216.  The  thickness,  marked  0,  of  the  pedestal  legs  in  Figs.  292,  29"),  and  280  is  not 
always  made  alike  by  the  different  locomotive  builders.      Our  practice  has  been   to 
make  the  thickness  o  for  straight  pedestal  legs  equal  to  the  depth  of  the  frame  brace 
B.,,  as  found  by  Rule  34«,  and  for  the  tapered  pedestal  legs,  the  thickness  o  in  the 
center  of  the  length  of  the  leg  was  also  made  the  same  depth.     These  dimensions  of 
the  pedestal  legs  have  always  given  good  satisfaction,  and  we  believe  can  be  safely 
adopted. 

217.  It  will  be  noticed  that  when  the  cast-iron  thimble  T  and  the  bolt  D  at  the 
bottom  of  the  pedestal,  Fig.  295,  are  used,  we  are  compelled  to  place  the  lower  frame 
brace  L  nearer  in  line  with  the  center  of  the  driving  axles  than  when  pedestal  eaps 
are  adopted,  as  shown  in  Fig.  292 ;  and  therefore  the  lower  frame  brace  L  in  Fig.  295 
will  be  subjected  to  a  greater  pulling  force  than  that  in  Fig.  292.     Hence,  for  finding 
the  depth  of  the  lower  frame  brace  L,  we  have  the  following  rules : 


194  MODERN   LOCOMOTIVE   CONSTRUCTION. 

BULE  35a. — When  east-iron  thimbles  at  the  bottom  of  the  pedestal  are  used,  as 
shown  in  Fig.  295,  multiply  the  depth  of  the  upper  frame  brace  B2,  as  found  by  Eule 
34a,  by  the  decimal  .86 ;  the  product  will  be  the  depth  of  the  lower  frame  brace  L. 

EULE  36. — When  pedestal  caps  are  used,  as  shown  in  Fig.  292,  multiply  the  depth 
of  the  upper  frame  brace  B2,  as  found  by  Eule  3-ia,  by  the  decimal  .69 ;  the  product 
will  be  the  depth  of  the  lower  frame  brace  L. 

EXAMPLE  66. — What  should  be  the  depth  of  the  lower  frame  brace  for  a  locomo- 
tive having  cylinders  18  inches  in  diameter,  when  cast-iron  thimbles  are  to  be  used  at 
the  bottom  of  the  pedestal,  and  the  maximum  steam  pressure  is  to  be  120  pounds  per 
square  inch  of  piston  I 

We  find  in  Fig.  271  that  the  depth  of  the  upper  frame  brace  B2,  suitable  for  this 
size  cylinder  and  steam  pressure,  is  3f  inches ;  hence  we  have,  according  to  Eule  35a : 

3.75  x  .86  =  3.22  inches  for  the  depth  of  the  lower  frame  brace  L. 

EXAMPLE  67. — What  should  be  the  depth  of  the  lower  frame  brace  for  a  locomo- 
tive having  cylinders  11  inches  in  diameter,  when  pedestal  caps  are  to  be  used  at  the 
bottom  of  the  pedestal,  and  the  maximum  steam  pressure  is  to  be  120  pounds  per 
square  inch  of  piston  ? 

In  Fig.  278  we  find  that  the  suitable  depth  of  the  upper  frame  brace  for  this  size 
cylinder  and  steam  pressure  is  2  inches ;  hence  we  have,  according  to  Eule  36 : 

2  x  .69  =  1.38  inch  for  the  depth  of  the  lower  frame  brace  L. 

Since  the  bottom  surface  of  this  brace  is  not  planed  along  the  entire  length,  that 
part  of  the  same  brace  to  which  the  pedestal  cap  is  bolted  is  usually  made  £  to  \ 
inch  deeper  than  the  depth  found  by  the  rule.  This  extra  depth  of  the  lower  frame 
brace  will  restore  some  of  the  strength  lost  by  the  holes  drilled  for  the  pedestal  cap 
bolts. 

THICKNESS   OF  THE  PEDESTAL  CAP. 

218.  The  thickness  C  of  central  portion  of  the  pedestal  cap,  Fig.  280,  is  usually 
made  |  inch  less  than  the  depth  of  the  lower  frame  brace  L. 

The  projections  M,  Fig.  280,  of  the  pedestal  jaw  generally  extend  into  the  cap  £  of 
an  inch  for  the  smaller  engines,  and  1  inch  for  the  larger  engines ;  and  since  the 
bottom  of  the  projections  u  are  generally  in  line  with  the  top  of  the  central  portion  C  of 
the  cap,  it  follows  that  the  ends  of  the  pedestal  cap  must  be  made  that  much  thicker. 

The  projections  u  are  slightly  tapered,  so  that  they  can  be  easily  entered  into  the 
recesses  in  the  cap,  and  when  the  cap  is  screwed  fast  into  position  they  will  firmly 
hold  the  ends  of  the  pedestal  jaw. 

NUMBER  OF  BOLTS   IN  PEDESTAL  CAPS. 

219.  It  is  the  general  practice  to  secure  the  pedestal  caps  in  small  engines  with 
two  bolts,  and  in  larger  engines  with  four  bolts.     We  believe  that  it  is  good  prac- 
tice to  use   two  bolts  for  each  pedestal   cap   in   locomotives   having  cylinders  14 
inches  and  less   in  diameter.     In   larger   locomotives  four  bolts  should  be  used  for 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


each  pedestal  cap.  The  di- 
ameters of  these  bolts  are 
given  in  Figs.  277,  278, 
279,  and  280. 

Fig.  292  represents  a 
frame  for  a  consolidation 
engine,    having   cylinders 
20  inches  in  diameter,  with 
all  the    pedestal  legs    ta- 
pered.  Formerly  nearly  all 
frames    had    pedestals  of 
this  kind,  but  lately  a  great 
number,  if  not  the  major- 
ity, of  frames  have  pedes- 
tals  in   which   one  leg  is 
straight,  as  shown  in  Fig. 
294.     It  will    be    noticed 
that    all    these    pedestals 
have  caps.     Fig.  295  rep- 
resents a  frame    for  the 
same  class  and  size  of  en- 
gine in  which  cast-iron  thimbles  are  used  at 
the  bottom  of  the  pedestal  jaw.    We  have 
frequently   seen   frames  with   this  kind  of 
pedestal  in  Mogul,  ten-wheeled,  and   eight- 
wheeled  engines,  but  have  not  seen  them 
used   in   frames   for   consolidation  engines. 
This  frame  has  been  designed  for  this  book, 
and  we  believe  it  to  possess  advantages  of 
its  own  for  consolidation  engines. 

BUILT-UP  FKAMES. 

220.  Fig.  297  represents  a  frame  which 
may  be  called  the  built-up  frame,  because, 
instead  of  it  being  forged  in  one  piece,  the 
same  as  all  the  frames  previously  shown, 
the  lower  brace  L  is  fitted  between  the  ped- 
estals and  bolted  to  the  same.  This  class  of 
frames  is  looked  upon  with  favor  by  a 
number  of  master-mechanics,  and  is  used 
to  a  comparatively  small  extent.  In  our 
opinion  the  built-up  frame  is  not  as  good 
as  either  one  of  the  solid  frames  shown  in 
Figs.  292  and  295;  it  lacks  that  simplicity 


^  -  ,. 


196  MODERN  LOCOMOTIVE   CONSTRUCTION. 

which  is  so  desirable  and  essential  in  a  locomotive.  The  frame  represented  in  Fig. 
297  is  one  of  the  frames  used  in  a  number  of  consolidation  engines  having  cylin- 
ders 20  inches  in  diameter.  We  consider  this  frame  to  be  too  light  for  engines  with 
cylinders  of  this  size. 

LIGHT   FRAMES. 

221.  It  has  been  found  that  when  the  depth  B2  of  the  upper  frame  brace  is  the 
same  throughout,  as  shown  in  Figs.  292  and  295,  and  when  the  brace  is  rather  light 
for  the  forces  which  it  has  to  resist,  fracture  will  take  place  somewhere  in  the  neigh- 
borhood marked  B3  in  Fig.  297,  near  the  pedestal,  and  seldom  midway  between  the 
pedestals.     Consequently,  when  it  is  necessary  to  make  the  frames  as  light  as  possible 
to  comply  with  given  conditions  of  a  railroad,  or  when  a  load  must  be  hauled  with  a 
locomotive  of  minimum  weight,  as  on  elevated  railroads,  the  weight  of  the  frames  can 
be  reduced  by  making  the  depth  B2  less  at  the  center  of  that  portion  of  the  upper 
frame  brace  which  connects  any  two  pedestals  without  weakening  the  frame.     When 
the  depth  of  the  brace  midway  between  the  pedestals  is  to  be  reduced,  then  make 
the  cross-sectional  area  and  the  depth  B3  near  the  pedestal  equal  to  that  found  accord- 
ing to  Rules  33a  and  34a ;  and  for  the  center  of  the  upper  frame  brace,  reduce  the 
depth  so  found  in  about  the  same  proportion  as  shown  in  Fig.  297. 

For  ordinary  locomotives  in  which  a  little  extra  weight  is  not  objectionable,  in 
fact  where  this  extra  weight  is  often  desirable,  it  is  always  best  to  leave  the  depth  of 
the  upper  frame  brace  the  same  throughout,  and  plane  the  under  side  of  the  brace. 
By  so  doing  an  advantage  will  be  gained  which  at  first  sight  may  appear  trivial,  but 
which  in  private  locomotive  shops  is  appreciated.  We  allude  to  the  bolts  which  are 
required  for  bolting  the  mechanism  to  the  upper  frame  braces.  These  bolts  are 
generally  made  before  they  are  actually  needed  in  the  erecting  shop,  and  according  to 
dimensions  obtained  from  the  drawing  room.  If  now  the  upper  frame  braces  are 
equal  in  depth  throughout  and  planed  to  correct  dimensions,  not  only  will  confusion, 
and  sometimes  the  necessity  of  throwing  bolts  away,  or  often  altering  the  lengths  of 
bolts,  be  avoided,  but  the  time  lost  by  the  workmen  waiting  for  bolts,  and  the  delay 
in  getting  the  engine  out  of  the  erecting  shop,  will  be  prevented,  which  otherwise 
would  have  amounted  to  quite  an  item  of  loss  to  the  proprietors. 

SLAB   FRAMES. 

222.  Sometimes  it  is  desirable,  and  particularly  so  in  narrow-gauge  locomotives, 
to  obtain  more  room  between  the  frames  for  the  fire-box  of  the  boiler  than  can  be 
obtained  by  leaving  the  frames  the  full  width  throughout.     In  cases  of  this  kind  the 
the  width  of  the  upper  frame  brace  B2  along  the  side  of  the  fire-box  is  reduced,  as 
shown  in  Fig.  299,  and  the  depth  of  the  brace  B.2  increased,  as  shown  in  Fig.  298.     In 
designing  locomotives  of  this  kind,  precautions  are  taken  to  bring  the  bottom  of  fire- 
box within  one  inch  from  the  top  of  the  lower  frame  brace  L,  and  never  allow  the 
bottom   of   the   fire-box   to   extend  below  this  brace;    by  so  doing  the  lower  frame 
brace  is  allowed  to  remain  the  full  width  of  the  frame,  and  room  is  also  provided  for 
the  spring  gear. 


MODEItX  LOCOMOTIVE   CONSTRUCTION.  197 

When  the  upper  frame  brace  is  to  be  made  in  the  form  of  a  slab,  as  shown,  its 
cross-sectional  area  for  any  given  diameter  of  cylinder  should  be  equal  to  that  of  the 
brace  suitable  for  the  same  diameter  of  cylinder,  and  found  according  to  Rule  33«. 
The  width  of  the  slab  is  arbitrary,  and  is  generally  made  as  small  as  is  prac- 
ticable in  the  designer's  ^__^_ 

judgment.  The  least  thick-     II       I  Fly.  29<J        ^  ^,  

711-ss  of  slab  that  we  have 
seen  was  li  inches,  used 
on  a  locomotive  having 
cylinders  15  inches  in  di- 
ameter: the  depth  of  the 


v,  n ,    •      u  Fig.  298 

same  brace  was  i  £  inches. 

From  the  foregoing  we  can  establish  the  following  rule  for  finding  the  depth  of 
the  frame  brace  /A,  when  it  is  to  be  of  the  slab  form: 

RULE  37. — First  find  the  cross-sectional  area  of  the  frame  brace  according  to 
Rule  33rt,  then  divide  this  area  by  the  given  width  of  the  slab ;  the  quotient  will  be 
the  depth  of  the  upper  frame  brace  or  slab. 

EXAMPLE  68. — What  should  be  the  depth  of  the  frame  brace  7?2  whose  width  is  1J 
inches  for  a  locomotive  having  cylinders  14  inches  in  diameter  ?  The  maximum  steam 
pressure  in  the  cylinder  is  to  be  120  pounds  per  square  inch. 

The  area  of  a  piston  14  inches  in  diameter  is  153.94  square  inches.  Hence,  accord- 
ing to  Rule  33a,  the  area  of  the  upper  frame  brace  will  be 

153.94  x  120 

— f)OOQ  '•  9.23+  square  inches. 

According  to  Rule  37,  the  depth  of  this  brace  will  be 

9  23 

^?  =  7.38,  say  7f  inches. 

FRONT   SPLICES   FOR  PASSENGER  LOCOMOTIVES. 

223.  The  general  design  of  the  front  splice,  sometimes  called  the  front  end  of 
the  frame,  depends  upon  the  class  of  locomotives  in  which  it  is  to  be  used.  Fig.  290 
represents  the  front  splice  for  an  eight-wheeled  passenger  locomotive.  The  manner 
of  fastening  the  front  splice  to  the  main  frame  depends  on  the  kind  of  pedestals 
adopted.  The  manner  of  fastening  the  splice  to  the  frame,  or,  we  may  call  it,  the  con- 
nection between  the  two,  when  pedestals  with  cast-iron  thimbles  and  bolts  are  used,  is 
shown  in  Fig.  289.  In  this  connection  the  keys  MM  are  usually  placed  in  a  position 
which  will  necessitate  the  drilling  out  a  small  portion  of  the  keys  so  as  to  allow  the 
bolts  N  N  to  pass  through  the  frame ;  this  will  prevent  the  keys  M  M  from  working 
out  of  position. 

In  Fig.  291  is  seen  the  manner  of  fastening  the  splice  to  frame  when  pedestals 
with  wrought-iron  caps  are  used.  In  this  connection  the  bolts  MM  which  fasten  the 
T-end  of  the  splice  to  the  pedestal  leg  are  liable  to  give  trouble  or  break;  to  prevent 
this,  great  care  must  be  taken  in  determining  the  diameters  of  these  bolts,  and  to  make 
them  as  large  in  diameter  as  possible  without  impairing  the  strength  of  the  pedestal 


198  MODERN  LOCOMOTIVE   CONSTRUCTION. 

leg.  To  determine  the  diameter  of  these  bolts,  we  have  the  following  rule,  which  is 
based  upon  observation  : 

RULE  38.  —  Multiply  the  width  of  the  frame  in  inches  by  the  decimal  .32  ;  the 
product  will  be  the  diameter  of  the  bolt  in  inches  for  fastening  the  T-end  of  splice  to 
pedestal  leg. 

EXAMPLE  69.  —  What  should  be  the  diameter  of  the  bolts  for  fastening  the  T-end 
of  splice  to  pedestal  leg  ?  The  width  of  frame  is  4  inches. 

4  x  .32  =  1.28,  say  1J  inches. 

These  bolts  have  usually  conical  heads,  and  countersunk  into  the  pedestal  leg. 

In  the  connection  of  splice  to  frame,  it  is  also  of  great  importance  to  have  a 
sufficient  number  of  bolts  N  N  to  hold  the  end  of  the  frame  to  splice.  The  diameter 
of  these  bolts  should  be  equal  to  about  J  of  the  width  of  the  frame,  and  the  shearing 
stress  should  not  exceed  3,000  pounds  per  square  inch.  Assuming  as  before,  for  the 
sake  of  simplicity,  that  the  total  pulling  force  is  equal  to  the  total  steam  pressure  on 
the  piston,  we  can  use,  for  determining  the  number  of  bolts  through  frame  and  splice, 
the  following  rule  : 

RULE  39.  —  Divide  the  total  steam  pressure  on  the  piston  hi  pounds  by  6,000  ;  the 
quotient  will  be  the  total  cross-sectional  area  of  all  the  bolts  ;  divide  this  quotient  or 
total  cross-sectional  area  by  the  cross-sectional  area  of  one  bolt  ;  the  quotient  will  be 
the  number  of  bolts  required  through  the  end  of  frame  and  splice. 

NOTE.  —  The  reason  for  dividing  the  total  steam  pressure  on  piston  by  6,000 
instead  of  3,000  is,  that  some  of  these  bolts,  frequently  all,  are  subjected  to  a  double 
shear  ;  that  is  to  say,  they  must  be  sheared  off  in  two  places  before  the  frames  can  be 
pulled  apart. 

EXAMPLE  70.  —  What  should  be  the  number  of  bolts  marked  N  N  in  Fig.  289, 
passing  through  the  end  of  frame  and  splice,  for  a  locomotive  having  cylinders  18 
inches  in  diameter,  maximum  steam  pressure  in  cylinder  120  pounds  per  square  inch  I 
The  diameter  of  each  bolt  to  be  equal  to  J  the  width  of  the  frame. 

In  Fig.  271  we  see  that  the  suitable  width  of  frame  for  a  locomotive  having 
cylinders  18  inches  in  diameter  is  4  inches  ;  hence  the  diameter  of  each  bolt  must  be  1 
inch.  The  cross-sectional  area  of  a  bolt  1  inch  in  diameter  is  .7854  of  a  square  inch. 

The  area  of  a  piston  18  inches  in  diameter  is  254.47  square  inches  ;  hence,  accord- 

254.47  x  120 
ing  to  Rule  39,  we  have,  --  (\rjfv\  —  "  =  5-089  square  inches  =  total  cross-sectional 

5.089 
area  of  all  the  bolts  ;  and  7Q~  (  =  6.4+  say  7  =  the  number  of  bolts  required. 


If  the  connection  of  frame  and  splice  is  similar  to  that  shown  in  Fig.  289,  and  the 
number  of  bolts  found  according  to  Rule  39,  and  also  assuming  that  the  total  pulling 
force  is  equal  to  total  steam  pressure  on  the  piston,  the  shearing  stress  per  square 
inch  of  cross-sectional  area  of  the  bolts  will  appear  to  be  greater  than  3,000  pounds, 
because  four  of  the  bolts  N  N  are  subjected  to  a  shear  in  one  place  only  ;  but  the  keys 
M  M  will  reduce  the  shearing  stress  to  less  than  3,000  pounds  per  square  inch  on  the 
bolts,  so  that  the  foregoing  rule  can  be  safely  applied  in  designs  of  this  kind. 

EXAMPLE  71.  —  What  should  be  the  number  of  bolts  N  N,  Fig.  291,  passing  through 


MODERN  LOCOMOTIVE   CONSTRUCTION.  199 

the  end  of  the  frame  and  splice  for  a  locomotive  having  cylinders  14  inches  in  diameter, 
tlif  diameter  of  each  bolt  is  to  be  equal  to  J  of  the  width  of  the  frame?  The  maxi- 
mum steam  pressure  in  cylinder  is  120  pounds  per  square  inch. 

In  Fig.  275  we  see  that  the  suitable  width  of  frame  for  a  cylinder  14  inches  in 

3jf 
diameter  is  3 $  inches,  consequently  the  diameter  of  each  bolt  N  N  must  be  -j  =  -f$ 

inch.  The  area  of  a  piston  14  inches  in  diameter  is  153.94  square  inches;  hence, 
according  to  Rule  39,  we  have, 

153.94  x  120 

-  =  3.0788  square  inches  for  the  total  cross-sectional  area  of  all  the  bolts. 

The  area  of  a  bolt  if  inch  in  diameter  is  equal  to  .69  of  a  square  inch,  and 

3.0788 


.69 


=  4.4+  say  5  bolts. 


224.  The  recess  marked  72,  Fig.  290,  near  the  front  end  of  the  frame  splice,  is  for 
the  purpose  of  receiving  the  cylinder  saddle,  which  generally  butts  against  the  rear 
end  of  the  recess.  The  cylinder  saddle  is  bolted  to  the  front  splice,  as  shown  in  Fig.  12, 
page  21,  by  bolts  running  in  a  horizontal  direction  through  the  flange  of  saddle  and 
splice,  and  also  by  the  vertical  bolts  B.  In  order  to  provide  further  security  and  pre- 
vent the  cylinder  from  moving  in  a  longitudinal  direction — that  is,  the  direction  in 
which  acts  the  greatest  force  which  the  cylinders  have  to  resist — a  key  D  is  driven  be- 
tween the  front  face  of  the  cylinder  saddle  and  end  of  recess.  Occasionally  we  find 
master-mechanics  using  two  keys  in  each  frame,  one  at  the  front  face  of  saddle  and 
another  one  at  the  rear  face.  We  prefer  to  use  only  one  key  at  the  front ;  and  believe 
this  to  be  the  best  practice,  because  two  frames  (sometimes  four)  are  usually  slotted  at 
one  time,  and  consequently  the  distances  in  the  frames  between  the  pedestals  and 
recesses  will  be  exactly  alike ;  the  facing  strips  on  the  cylinder  saddles  are  planed  in 
line  and  square  with  the  axis  of  cylinders,  and  therefore  by  placing  the  cylinder  saddle 
directly  against  the  rear  ends  of  the  recesses,  the  cylinders  are  brought  in  the  true 
position  with  less  labor  than  when  two  keys  are  used  in  each  frame. 

'!'!'•>.  In  passenger  engines,  the  lifting-shaft  bearing  and  rocker-box,  besides 
other  mechanism,  are  bolted  to  the  front  splice,  consequently  it  is  subjected  to  the 
action  of  vertical  forces  of  considerable  magnitude,  and  it  has  also  to  resist  the  pulling 
force  due  to  the  pressure  on  one  piston ;  therefore  in  this  class  of  locomotives  it  is 
generally  made  somewhat  deeper  than  the  upper  frame  brace  B2  in  Fig.  289,  but 
uniformity  in  the  proportion  of  these  depths  does  not  exist.  As  a  result  of  observa- 
tion on  this  point,  we  believe  that  the  following  rule  will  give  a  depth  for  the  front 
splice  which  will  agree  with  good  modern  practice. 

RULE  40.— Multiply  the  depth  of  the  upper  frame  brace  #>,  Fig.  289,  by  1.15 ;  the 
product  will  be  the  depth  of  the  front  splice. 

According  to  this  rale  the  depth  of  the  front  splice  is  15  per  cent,  deeper  than  that 
of  the  upper  frame  brace. 

NOTE. — When  the  maximum  steam  pressure  on  the  piston  is  120  pounds  per  square 
inch,  then  take  the  depth  of  the  upper  frame  brace  B.,  from  a  pedestal,  suitable  for  the 
given  diameter  of  cylinder,  shown  in  the  group  Figs.  271  to  l27!>. 


200 


MODEEN  LOCOMOTITE   CONSTRUCTION. 


When  the  maximum  pressure  is  more  or  less  than  120  pounds  per  square  inch  of 
piston,  then  find  the  depth  of  the  upper  frame  brace  B2  by  Eules  33«  and  34a. 

EXAMPLE  71a. — What  should  be  the  depth  of  the  front  splice  for  a  locomotive 
having  cylinders  18  inches  in  diameter?  Maximum  steam  pressure  on  pistons  is  120 
pounds  per  square  inch. 

In  Fig.  271  we  see  that  the  depth  of  the  upper  frame  brace  B.,  for  an  18-inch 
cylinder  is  3|  inches ;  hence,  according  to  Rule  40,  we  have, 

3.75  x  1.15  =  4.3125  inches, 

which  is  the  depth  of  the  front  splice.     The  depth  of  the  front  splice  at  the  recess 
should  be  equal  to  the  depth  of  that  portion  of  the  upper  frame  brace  which  is  marked 

B2  (see  Fig.  289),  and  the  depth  of  the 
splice  from  the  cylinders  to  the  bump- 
ers should  be  equal  to  the  depth  B%. 

226.  The  front  end  P  of  the  splice 
is  often  turned  down,  forming  an  off- 
set, as  shown  in  Fig.  290.  To  this 
offset  is  bolted  the  bumper,  usually 
made  of  wood.  In  some  locomotives 
the  offset  at  P  is  in  an  upward  direc- 
tion, so  as  to  bring  the  bumper  to  a 
suitable  height  above  the  rails  for  con- 
venience in  coupling  to  the  cars. 

In  cases  of  accidents  or  collision, 
the  front  end  P  of  the  splice  is  very 
liable  to  be  broken  off  or  otherwise 
injured,  and  to  repair  the  damage  the 
whole  splice  will  have  to  be  taken  off. 
To  obviate  this  difficulty  and  thus  save 
considerable  time  and  labor,  many 
master-mechanics  now  make  the  splice 
perfectly  straight  at  the  front  end,  and 
in  place  of  the  offset  P  use  a  cast- 
ing, of  which  an  elevation  and  plan  are 
shown  in  Figs.  300  and  301.  This  cast- 
ing here  shown  is  suitable  for  loco- 
motives having  cylinders  17  inches  in 
diameter ;  for  smaller  engines  the  dimensions  may  be  somewhat  decreased,  and  for 
larger  ones  they  should  be  increased. 

227.  Figs.  293  and  296  represent  the  front  splices  for  consolidation  engines,  and  a 
similar  form  of  splice  is  also  often  used  for  Mogul  engines.  These  splices  pass  over  the 
top  of  cylinder  saddle,  and  are  fastened  by  bolts  D  I),  Figs.  292  and  295,  passing  through 
the  front  splice,  cylinder  saddle,  and  front  end  of  main  frame ;  they  are  also  fastened 
to  the  main  frame  by  the  bolts  N  N,  and  further  secured  in  position  by  the  keys  R  7?, 
and  also  by  the  keys  K  K  between  the  cylinder  saddle  and  end  of  recess  in  the  splice. 


MODEBX   LOCOMOTIVE    CONSTRUCTION. 


201 


DEPTH   OF   FRAME   SPLICES   FOR  MOGUL   AND   CONSOLIDATION   ENGINES. 

228.  When  the  form  of  the  splice  is  like  that  shown  in  Fig.  295,  the  depth  of  that 
part  of  the  splice  marked  S  between  the  main  frame  and  saddle,  and  also  that  portion 
of  the  splice  which  lies  on  top  of  cylinder  saddle,  should  be  equal  to  that  of  the  upper 
frame  brace  marked  It.,.     The  depth  of  the  splice  in  front  of  the  cylinder  saddle  can  be 
made  equal  to  that  of  the  upper  frame  brace  which  is  marked  B3. 

FRONT  SPLICE  AND  MAIN   FRAME  FORGED   IN   ONE  PIECE. 

229.  Sometimes  the  front  splice  and  main  frame  are  forged  in  one  piece,  as  shown 
in  Fig.  302.     This  we  consider  to  be  very  bad  practice,  because,  should  the  front  end 


Fig.  303 


be  injured,  the  whole  frame  will  have  to  be  taken  down  to  repair  the  damage,  which 
will  involve  a  great  amount  of  unnecessary  labor  and  expense. 

FRAME  BOLTS. 

230.  The  bolts  which  are  used  for  bolting  the  different  parts  of  the  engine  to  the 
frames  are,  by  the  majority  of  locomotive  builders,  made  straight,  accurately  turned 
to  fit  reamed  holes,  and  driven  home.  Other  builders  make  these  bolts  tapered, 
generally  i  inch  to  the  foot — that  is,  in  the  length  of  one  foot  the  diameter  is  increased 
|  inch — and  turn  them  to  such  dimensions  as  will  allow  the  bolts  to  enter  the 
reamed  holes  to  within  J  of  an  inch  from  the  head  of  the  bolt ;  through  this  distance 
of  i  inch  the  bolts  are  driven  home.  We  are  inclined  to  believe  that  the  use  of 
tapered  bolts  is  the  best  practice,  as  when  these  bolts  become  slack,  then  by  turning  a 
small  amount  off  the  under  side  of  head,  the  bolts  can  again  be  driven  tightly  into  the 
holes.  We  also  believe  that  tapered  bolts  will  hold  the  parts  more  firmly  together  than 
straight  bolts.  Bolts  which  are  very  long,  as  those  marked  I)  />  in  Figs.  2!>l>  and  295, 
should  always  be  tapered,  which  we  believe  to  be  the  best  practice. 


SI 'I, 'ING    SADDLES. 


231.  Fig.  303  represents  the  pedestal  /',  with  driving  axle  box  /?,  wodire.x  IT  W2, 
and  spring  saddle  .V  in  position.  A  portion  of  the  spring  /and  also  the  spring  strap 
H  are  shown.  Fig.  304  represents  a  vertical  section  through  the  centers  of  the  same 


202 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


mechanism.  Spring  saddles  of  the  form  here  shown  are  made  of  cast-iron,  and  are 
suitable  for  locomotives  having  cylinders  larger  than  12  inches  in  diameter.  The  base 
E  E  of  these  saddles  is  made  as  wide  as  the  space  between  the  wedges  will  permit, 
leaving  a  sufficient  amount  of  metal  around  the  recesses  in  top  of  the  driving  box  in 
which  the  base  of  the  saddle  is  placed.  Pai't  of  the  metal  at  F,  in  the  base  E  E  of  the 


Jfty.  SOS 


i 

-W- 


Fio-\30G  Fig.  307 

ffroiiffltt  Iron  Saddle 


saddle,  is  cut  out,  thereby  giving  access  for  oiling  the  axle  journal.  A  saddle  with  this 
kind  of  base  will  have  a  firmer  support  than,  and  is  not  so  liable  to  upset,  as  a  saddle 
with  a  narrow  base,  similar  in  form  to  that  shown  in  Fig.  306. 

The  width  of  the  spring  saddle  is  made  to  allow  at  least  J  inch  clearance  at 
each  side  of  the  frame,  as  shown  in  Fig.  304.  A  recess  is  cast  in  the  top  of  the  saddle 
to  receive  the  spring  strap  H.  At  the  center,  in  the  bottom  of  the  recess,  a  fulcrum  G 
is  cast  to  fit  into  a  groove  cut  in  the  bottom  of  the  spring  strap  H,  the  whole  arrange- 
ment being  such  as  will  allow  the  spring  I  to  rock  through  a  short  distance. 

232.  Some  locomotive  builders  do  not  cast  the  fulcrum  G  in  the  recess,  but  make 
the  bottom  of  it  perfectly  flat,  as  shown  in  Fig.  305.     In  this  recess  a  piece  of  rub- 
ber E  is  placed,  with  a  wrought-iron  plate  P  on  top.     This  plate  is  about  i  inch 
thicker  in  the  center  than  at  the  ends,  so  that  when  the  spring  strap  (in  this  case 
without  a  groove  cut  in  it)  is  placed  on  top  of  the  plate,  the  spring  can  rock  the 
required  amount.     In  this  kind  of   spring  saddle  the  recess  should  always  be  deep 
enough  to  allow  the  spring  strap  to  enter  it  at  least  J  inch,  to  prevent  the  spring  from 
moving  out  of  position. 

233.  Sometimes  the  spring  saddles  are  made  of  wrought-iron,  their  form  being 
similar  to  that  of  the  cast-iron  ones,  with  the  exception  of  the  recess,  which  is  left  off. 
The  tops  are  straight,  and  a  pin  A,  as  shown  in  Fig.  306,  screwed  into  the  top.     This 
pin  is  made  to  fit  loosely  in  a  hole  in  the  bottom  of  the  spring  strap,  and  prevents  the 
spring  from  moving  out  of  position.     The  bottom  of  the  strap  is  made  convex,  to 
allow  the  spring  to  rock. 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


203 


Occasionally,  when  wrouglit-iron  spring  saddles  are  used,  a  roller  is  inserted 
between  the  saddle  and  strap,  lying  in  suitable  grooves  cut  into  both.  The  roller  has 
also  collars  at  each  end,  to  prevent  the  spring  from  slipping  sideways. 

For  locomotives  having  cylinders  12  inches  in  diameter  and  less,  the  straps  are 
made  of  wrought-iron,  of  a  form  as  shown  in  Figs.  306  and  307,  which  need  no 
further  explanation. 

DRIVING  AXLE  BOXES. 

234.  The  play  or  clearance  between  the  axle  box  and  hub  of  driving  wheel  F,  Fig. 
304,  is  generally  ?-6  inch,  and  the  clearance  between  the  axle  box  and  collar  U  is 
equal  to  the  same  amount,  giving  the  driving  axle  a  total  amount  of  £  inch  play  or 
movement  in  the  direction  of  its  length. 

Occasionally  we  find  driving  axle  boxes  in  which  the  distance  K  (Fig.  304),  that 
is,  the  distance  from  the  center  line  X  Y  to  the  outside  face  of  box,  is  less  than  the 
distance  L  from  the  center  line  X  Y  to  the  inner  face  of  box,  the  difference  generally 


Fig.  313 


being  J  to  ,-%  inch.  The  object  of  this  difference  is  for  the  purpose  of  turning  the  axle 
box  around  to  take  up  the  wear  between  it  and  the  face  of  wheel,  when  that  becomes 
excessive. 

235.  Driving  axle  boxes,  or  which,  for  the  sake  of  brevity,  are  often  called  driving 
boxes,  consist  essentially  of  three  parts,  namely,  the  casting  marked  A  A  in  Fig.  308, 
an  oil  cellar  C,  and  the  brass  B.  The  driving  boxes  may  be  divided  into  two  classes, 
the  form  of  the  brass  B  being  the  distinguishing  feature. 

Figs.  308,  309,  310  represent  different  views  of  one  class  of  driving  boxes,  in 


204  MODERN  LOCOMOTIVE   CONSTRUCTION. 

which  the  octagonal  fonn  of  brasses  are  used,  and  Figs.  311,  312,  313  represent  the 
other  class  of  driving  boxes,  in  which  the  cylindrical  form  of  brasses  are  used. 

Occasionally  we  find  the  casting  A  A  made  of  brass;  in  such  cases  a  separate 
brass  bearing  is  not  used ;  but  boxes  of  this  kind  are  seldom  adopted,  and  therefore  we 
will  confine  our  attention  to  that  class  of  boxes  which  are  called  cast-iron  driving  boxes. 

In  proportioning  these  boxes  great  care  must  be  taken  to 'have  the  depth  f/f  of 
the  lug  A2,  Fig.  311,  sufficiently  great  so  that  it  cannot  be  broken  off  by  pressing  the 
brass  into  the  box,  or  the  strength  of  the  lug  A2  impaired,  causing  it  to  break  off 
when  the  engine  is  running.  In  large  driving  boxes  the  depth  g  f  should  be  at  least 
1J  inches. 

In  a  great  number  of  locomotives  the  inner  faces  of  the  flanges  m  m,  Fig.  312,  are 
planed  parallel  to  each  other,  and  in  a  small  number  of  engines  the  inner  surfaces  of 
these  flanges  are  planed  to  a  form,  as  shown  in  Fig.  313;  that  is,  when  planed  the 
flanges  are  about  -^  inch  thicker  in  the  center  p  than  at  the  ends  o  o.  The  object 
of  this  form  of  driving-box  flange  is  to  prevent  the  same  from  binding  against  the 
pedestal  wedge,  and  at  the  same  time  give  the  box  a  greater  freedom  to  adjust  itself 
to  the  journal  when  one  end  of  the  axle  stands  higher  than  the  other  end,  caused  by 
running  over  an  uneven  track. 

The  width  I  I  of.  the  flanges,  Fig.  312,  should  be  sufficient  to  allow  their  lower 
ends,  when  the  box  is  in  the  lowest  position  in  the  pedestal,  to  cover  the  pedestal  legs. 
By  this  arrangement  the  lateral  stress  on  the  flanges  of  the  wedges  will  be  less  than 
when  the  driving-box  flanges  do  not  reach  the  pedestal  legs. 

236.  The  oil  cellar  (7,  Fig.  311,  is  made  of  cast-iron,  and  its  purpose  is  to  hold  the 
waste,  tallow,  and  oil  to  lubricate  the  journal.     The  ends  of  the  cellar  are  planed  to 
fit  accurately  in  the  box ;  and  it  is  held  in  the  box  by  two  bolts  r  r,  which  are  roughly 
turned  and  fit  in  the  holes  somewhat  loosely ;  these  bolts  are  secured  in  position  by 
means  of  the  split  keys  s.     In  many  driving  boxes  the  end  surfaces  u  u  of  the  cellar 
are  pai'allel  to  each  other ;  in  others  we  find  these  surfaces  inclined  towards  the  cen- 
ter, making  the  width  at  the  top  of  the  cellar  about  £  inch  less  than  at  the  bottom. 
The  reason  for  tapering  the  width  of  the  cellar  is,  that  it  must  be  removed  before 
the  driving  box  can  be  taken  off  the   axle;   but   experience   has   shown   that  the 
lower  ends  of  the  driving  box  will  close  after  it  has  been  in  use  for  some  time,  and 
clasp  the  oil  cellar  very  tightly,  and  therefore  a  tapered  oil  cellar  can  be  more  readily 
removed  when  it  is  necessary  to  do  so  than  one  with  parallel  ends. 

237.  The  pockets  n  «,  Figs.  311  and  312,  in  the  top  of  the  driving  box,  receive 
the  ends  of  the  spring  saddle  and  prevent  the  latter  from  moving  out  of  position. 

The  recesses  k  k  are  for  the  purpose  of  leading  all  the  oil  which  may  be  poured  on 
the  top  of  the  box  into  the  oil  holes  i  i. 

DRIVING-BOX  BRASSES. 

238.  There  are  some  objections   to  the  use  of  octagonal  brasses ;    for  instance, 
they  require  a  considerable  amount  of  labor  to  fit  them  in  the  box  as  accurately  as  they 
should  be ;  and  again,  since  the  ends  of  these  brasses  are  not  firmly  secured  in  the 
box,  they  are  liable  to  close,  press  against  the  axle  journal,  and  consequently  become 


MonKii\  LocoMOTirs  coxsTnrcTiny.  205 

hot  in  a  very  short  time,  and  for  these  reasons  the  octagonal  form  of  brass  is  not 
extensively  used. 

The  eylindrieal  form  of  brass  shown  in  Fig.  311  gives  better  satisfaction,  and  is 
the  form  adopted  in  a  large  majority  of  locomotives.  Its  outer  surface  is  accurately 
turned,  the  casting  A  A  is  slotted,  and  the  brass  pressed  in  with  a  pressure  of  about 
five  to  seven  tons.  The  edge  d  of  all  driving-box  brasses  generally  extends  i  to  f 
inch  below  the  center  of  the  axle;  and  the  edge  g  generally  extends  £  inch  below 
the  edge  d;  the  object  of  this  form  is  to  hold  the  ends  securely  in  position,  so 
as  to  prevent  them  from  closing,  and  thereby  avoid  hot  journals.  Even  when  these 
brasses  are  accurately  fitted  and  pressed  in  the  box  very  tightly,  they  will  in  time 
become  loose,  and  will  have  to  be  replaced.  In  order  to  make  these  brasses  remain 
tight  in  the  box  for  a  greater  period  of  time,  some  master-mechanics  will  drive  two 
brass  pins  t  t,  about  f  inch  in  diameter,  through  each  side  of  the  box  and  brass. 
These  pins  are,  for  obvious  reasons,  driven  at  an  angle  with  the  sides  of  the  box,  as 
shown  in  Fig.  311.  Another  advantage  gained  by  these  pins  is  that,  when  collars 
on  the  driving  axle  are  not  used,  as  is  sometimes  the  case,  the  brasses,  when  they 
have  become  loose,  cannot  slip  out  of  the  box. 

In  the  top  of  the  brass  is  cast  an  oil  groove  h,  about  £  inch  square.  This  groove 
extends  to  within  1  inch  from  the  ends  of  the  brass.  The  oil  is  led  into  this  groove 
by  two  oil  holes  i  i,  each  about  £  inch  in  diameter. 

Babbitt  metal  is  used  in  many  driving-box  brasses.  Grooves,  about  J  to  1  inch  in 
width,  and  extending  sometimes  the  whole  length  of  the  brass,  and  at  other  times  to 
about  3  inch  from  the  end  of  the  same,  are  cast  into  and  near  the  top  of  the  brass, 
as  indicated  by  </  <j  in  Figs.  308  and  310. 

We  believe  that  the  Babbitt  metal  in  driving-box  brasses  is  worse  than  useless, 
because  the  waste  in  the  oil  cellars  will  accomplish  the  same  purpose  for  which  Babbitt 
metal  is  intended,  namely,  to  prevent  the  dust  from  spreading  around  the  axle  journal. 
Besides,  in  our  experience,  we  have  found  that,  since  the  pressure  of  the  brass  against 
the  journal  is  very  great  and  acting  constantly,  and  since  the  Babbitt  metal  will  collect 
and  hold  the  dust,  the  axle  journal  will  wear  comparatively  very  rapidly,  and  for  these 
reasons  the  brass  is  better  without  it. 


PROPORTIONS   OF  DRIVING  AXLE  BOXES. 

239.  In  designing  a  locomotive  driving  box  we  must  not  lose  sight  of  the  fact  that 
all  the  weight  which  is  placed  upon  it  must  be  supported  by  the  upper  part  of  the  axle 
journal.  Indeed,  herein  lies  a  great  difference  between  a  locomotive  driving  box  and 
an  ordinary  pillow  block  similar  to  those  used  in  many  stationary  engines.  In  the  for- 
mer the  pressure  is  against  the  upper  part  of  the  box,  and  consequently  the  oil  which 
is  fed  through  the  oil  holes  in  the  top  of  the  box  will  be  aided  to  flow  away  from  the 
elements  of  contact.  In  the  ordinary  pillow  block  the  pressure  is  against  its  lower  part, 
and  if  the  pressure  is  not  sufficiently  intense  to  force  out  the  lubricant  from  between 
the  surfaces,  the  oil  will  be  aided  to  some  extent  to  flow  towards  the  element  of 
contact. 

It  must  also  be  remembered  that  the  vertical  pressure  on  a  locomotive  axle  journal 


206  MODERN  LOCOMOTIVE   CONSTRUCTION. 

is  almost  constant,  which  will  in  nowise  assist  in  the  lubrication  of  the  journal.  In 
pillow  blocks  of  stationary  engines,  although  the  pressure  is  generally  towards  the  bot- 
tom of  the  block,  it  will  shift  a  little  from  one  side  of  the  pillow  block  to  the  other  as 
the  piston  changes  the  direction  of  its  motion,  and  thereby  assist  in  the  lubrication  of 
the  journal. 

These  considerations  lead  us  to  conclude  that  locomotive  axle  journals  are  more 
liable  to  become  hot  than  the  main  journals  of  stationary  engines.  To  prevent  as 
much  as  possible  hot  axle  journals,  we  must  proportion  them  in  a  manner  which  will 
allow  upon  them  a  comparatively  low  pressure  per  square  inch.  The  pressure  on  any 
journal  is  estimated  by  the  pressure  per  square  inch  of  its  projected  area.  By  the  pro- 
jected area  is  meant  the  area  of  a  rectangle  whose  length  and  breadth  are  respectively 
equal  to  the  length  and  diameter  of  the  journal. 

240.  The  pressure  on  an  axle  due  to  the  weight  of  the  engine  is  not  the  only  press- 
ure which  the  axle  journal  has  to  resist ;  it  has  also  to  resist  a  pressure  due  to  the  load 
which  the  engine  has  to  haul.     The  latter  pressure  is  not  a  constant  quantity,  but  the 
ratio  between  the  pressure  due  to  the  weight  of  the  engine  and  that  due  to  the  maxi- 
mum load  is  about  the  same  in  all  engines ;  and  therefore,  for  the  sake  of  simplicity, 
we  may  leave  the  pressure  due  to  load  out  of  the  question,  and  proportion  the  journal 
according  to  the  pressure  due  to  the  weight  of  the  engine. 

Close  observation  and  experience  in  modern  locomotive  construction  lead  us  to  be- 
lieve that  a  pressure  of  about  160  and  not  ovev  175  pounds  per  square  inch  of  projected 
area,  due  to  the  weight  of  the  engine,  agrees  with  the  best  modern  practice  and  may  be 
adopted ;  the  former  figure,  namely,  160  pounds  per  square  inch,  should  be  preferred ; 
it  will  give  the  best  results. 

241.  From  the  foregoing  it  will  be  seen  that,  in  order  to  design  the  driving  box, 
we  must  first  determine  the  total  weight  which  the  driving  axle  journals  will  have  to 
support. 

The  weight  on  the  driving  axle  journals  for  new  locomotives  can  generally  be  esti- 
mated only  approximately,  because  the  design  is  not  sufficiently  advanced  to  obtain 
the  exact  weights  of  the  different  parts  of  the  engine  and  running  gear ;  yet  experience 
will  enable  us  to  estimate  this  weight  close  enough  for  all  practical  purposes.  In  Art. 
24,  tables  will  be  found  which  give  the  weights  on  drivers  in  the  various  classes  of 
locomotives.  Table  5  contains  the  weights  on  drivers  in  eight- wheeled  passenger  en- 
gines, to  which  we  shall  refer  here.  Now,  in  order  to  determine  the  weight  which  the 
driving  axle  journals  in  eight-wheeled  passenger  engines  will  have  to  support,  we  must 
subtract  the  sum  of  the  weights  of  the  wheel  centers,  tires,  axles,  side  rod,  etc.,  in  fact, 
the  weight  of  all  pieces  which  are  supported  directly  by  the  track,  from  the  total  weight 
on  drivers  given  in  Table  5 ;  one-fourth  of  the  remainder  will  be  the  weight  on  each 
journal. 

If  the  design  of  the  engine  is  not  sufficiently  advanced  to  obtain  the  accurate 
weights  of  the  driving  wheels,  axles,  side  rods,  etc.,  we  can  generally  estimate  the  sum 
of  these  weights  by  assuming  them  to  be  from  £  to  \  of  the  total  weight  on  the  drivers. 
Thus,  let  it  be  required  to  find  the  weight  on  the  driving  axle  journals  for  an  eight- 
wheeled  passenger  engine  having  cylinders  16  inches  in  diameter.  In  Table  5  we  find 
the  total  weight  on  the  drivers  for  a  locomotive  of  this  size  to  be  47,665  pounds.  One- 


MODERN  LOCOMOTirE  CONSTRUCTION. 


I 

I 


Kl 


L X»  —  A KS ^ 

1  "  I       " 


I 


••^--4-^-r 


208 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


fourth  of  this  weight  will  be  the  estimated  sum  of  the  weights  of  the  parts  supported 

47665 


directly  by  the  track.      Hence,  47665  - 


weight  supported  by  all  the  driving  axle  journals;  and 


=  35749  pounds,  which   is  the   total 
35749 


=  8937  pounds  sup- 


(• — as — f~~t~' 


ported  by  each  driving  axle  journal ;  or,  we  may  say  that  the  pressure  on  the  projected 

area  of  the  journal  is  8,937  pounds ;  and  since  the 
pressure  per  square  inch  is  to  be  about  160  pounds 
and  not  exceed  175  pounds,  we  must  proportion  the 
diameter  and  length  of  the  journal  accordingly.  By 
the  term  "  length  of  journal "  we  mean  that  length 
which  is  equal  to  the  width  A  (Fig.  316)  of  the 
driving  box,  and  neglecting  the  short  length  of 
journal  required  for  the  play  between  the  axle  box 
and  hub  of  driving  wheel. 

242.  When  the  driving  axle  is  made  large  enough 
to  prevent  heating,  and,  to  fulfill  conditions  peculiar 
to  the  locomotive,  also  properly  fitted  into  the  hub 
of  the  wheel,  we  have  generally  an  axle  strong 
enough  to  resist  all  the  forces  which  may  act  upon 
it,  and  therefore  in  determining  the  diameter  and 
length  of  journal  wre  may  throw  out  of  considera- 
tion the  strength  of  an  axle. 

One  of  the  conditions  peculiar  to  the  locomotive, 
and  to  which  a  driving  axle  must  conform,  is  the 
width  A  of  the  axle  box  (Fig.  316),  which  is  short 
when  compared  with  the  lengths  of  pillow  blocks  in 
stationary  or  marine  engines.  This  width  A  of  the 
axle  box  is  limited  by  the  distance  between  the  cylin- 
ders and  also  by  the  gauge  of  the  track ;  the  latter  is 
established,  and  therefore  if  we  make  the  axle  box 
too  wide,  then,  since  the  center  of  the  width  of  frame 
and  the  center  of  axle  box  should. coincide,  or  very 
nearly  so,  we  must  either  move  the  frames  closer 
together  to  a  corresponding  amount,  and  thus  be 
compelled  to  reduce  the  width  of  the  fire-box,  which 
is  decidedly  an  objectionable  feature,  or  we  must 
increase  the  distance  between  the  cylinders,  which 
is  also  objectionable. 

The  width  A  of  the  axle  box  should  be  only  suf- 
ficient to  allow  for  the  proper  thickness  of  the  flanges 
on  the  box  and  wedges  to  give  these  the  requisite 
strength. 

Figs.  314  to  340  represent  driving  boxes  suit- 
able for  the  pedestals  shown  in  Figs.  271  to  279, 


LOCOMOTIVE   COXSTRTCTION.  209 

and  are  designed  for  eight-wheeled  passenger  locomotives  of  various  sizes,  includ- 
ing those  having  cylinders  10  inches  in  diameter,  and  -up  to  18  inches  in  diam- 
eter. The  given  dimensions  of  these  boxes  agree  with  the  average  modern  loco- 
motive practice.  From  those  illustrations  we  can  readily  obtain  the  width  of  the 
boxes,  and  all  the  main  dimensions.  The  part  shown  in  section  at  the  right-hand 
side  of  all  the  plans  of  the  boxes,  such  as  Figs.  316,  319,  322,  etc.,  and  marked  W 
in  some  of  them,  represent  a  section  of  the  long  wedge;  the  distance  between  the 
flanges  of  the  wedges  is  the  thickness  of  the  frame,  and  this  thickness  subtracted  from 
the  distance  between  tlje  flanges  of  the  boxes  and  divided  by  two  will  give  the  thick- 
ness of  each  flange  on  the  wedge. 

•J4:>.  Having  established  the  width  of  the  driving  boxes,  the  diameter  of  the  journal 
is  easily  obtained  by  the  following  rule  : 

RULE  41.  —  Divide  the  weight  on  the  journal  by  160  ;  the  product  will  be  the  num- 
ber of  square  inches  in  the  projected  area  ;  divide  this  quotient  by  the  width  A  (Figs. 
316  to  340)  of  the  driving  box  in  inches;  the  quotient  will  be  the  diameter  of  the 
journal  in  inches. 

EXAMPLE  72.  —  What  should  be  the  diameter  of  a  driving  axle  journal  for  an  eight- 
wheeled  passenger  engine  having  cylinders  16  inches  in  diameter  ? 

We  have  already  seen  in  Art.  241  that  the  total  pressure  on  the  projected  area 
of  one  driving  axle  journal  for  this  class  and  size  of  engine  is  8,937  pounds.  In  Fig. 
322  we  find  that  the  width  of  box  should  be  8  inches.  Hence  to  find  the  diameter  we 

8937 
have,  according  to  Eule  41,  -rr  =  55.8,  which  is  the  number  of  square  inches  in  the 


K.r.  o 

projected  area,  and  —^~  =  6.97  inches  =  diameter  of  the  journal.     If  we  make  the 

diameter  of  this  axle  equal  to  that  given  in  Fig.  320,  namely,  6J  inches,  the  pressure 
per  square  inch  on  the  projected  area  will  be  165.5  pounds,  providing  our  estimated 
weight  of  wheels,  tires,  axles,  etc.,  is  exactly  correct. 

EXAMPLE  73.  —  Find  the  diameter  of  a  driving  axle  suitable  for  an  eight-wheeled 
passenger  engine  whose  cylinders  are  12  inches  in  diameter. 

In  Table  5  we  find  the  total  weight  on  the  drivers  to  be  29,700  pounds.  Allowing 
|  of  the  total  weight  on  the  drivers  for  the  mechanism  whose  weight  is  not  supported 

29700 

by  the  journals,  we  have,  29700  -  -  =  23760  pounds  pressure  on   all  the  four 

o 

23760 
journals  and  —  ~r~   =  5940  pounds  pressure  on  the  projected  area  of   each  journal. 

5940 
Again,      ™  =  37.12  square  inches  in  each  projected  area.     In  Fig.  334  we  find  that 


for  a  passenger  locomotive  with  cylinders  12  inches  in  diameter,  the  width  of  the 
driving  box   should  be  6f   inches,  and  consequently  the  diameter  of  the  journal 

37  12 
should  be    fi  '„-    =  5.49  inches.      These   dimensions   agree  with  those  given  in  the 

illustrations. 

From  the  foregoing  it  will  bo  seen  that  in  determining  the  dimensions  of  a  locomo- 
tive driving  axle  journal,  we  have  followed  the  law  of  simple  proportionality  of  friction 


210  MODERN  LOCOMOTIVE   CONSTRUCTION. 

to  pressure.  In  relation  to  this  law,  Prof.  W.  J.  M.  Rankine  says:  The  law  of 
simple  proportionality  of  friction  to  pressure  is  only  true  for  dry  surfaces,  when  the 
pressure  is  not  sufficiently  intense  to  indent  or  grind  the  surfaces ;  and  for  greased 
surfaces,  when  the  pressure  is  not  sufficiently  intense  to  force  out  the  unguent  from 
between  the  surfaces  where  it  is  held  by  capillary  attraction.  If  the  proper  limit  of 
intensity  of  pressure  be  exceeded,  the  friction  increases  more  rapidly  than  in  the 
simple  ratio  of  the  pressure.  That  limit  diminishes  as  the  velocity  of  rubbing 
inci*eases,  according  to  some  law  not  yet  exactly  determined.  The  following  are  some 
of  its  values,  deduced  from  experience : 

RAILWAY    CARRIAGE   AXLES. 

Limit  of  pressure 
per  square  inch. 

Velocity  of  rubbing  surface,  1  foot  per  second 392 

"  "  "        2i  feet    "        "     224 

«  "  "          g       "        «          tt  i  AQ 

The  limit  of  the  pressure  on  journals  given  by  Professor  Rankine  is  exceeded  on 
the  journals  proportioned  by  the  foregoing  rules,  and  such  as  are  used  in  modern 
locomotives,  yet  these  journals  are  giving  good  results,  and  seem  to  be  suitable  for  the 
purpose  intended.  But  if  the  average  speed  of  the  present  locomotive  is  to  be 
increased,  so  that  the  velocity  of  the  circumference  of  the  journal  exceeds  9  feet  per 
second,  the  lengths  of  these  journals  may  also  have  to  be  increased,  so  as  to  reduce  the 
pressure  per  square  inch  to  considerably  less  than  160  pounds,  even  if  we  are  com- 
pelled to  increase  the  distance  between  the  cylinders. 

244.  The  driving  boxes  shown  in  Figs.  314  to  340  were  proportioned  to  suit  one 
particular  class  of  locomotives,  namely,  eight-wheeled  passenger  engines.  From  our 
remarks  in  Art.  243  it  will  be  seen  that  the  size  of  a  box  for  a  passenger  engine  depends 
upon  the  weight  on  the  drivers ;  and  since  the  cylinders  are  proportioned  in  accord- 
ance with  the  weight  on  the  drivers,  we  may  assume  and  say,  as  our  illustrations 
indicate,  that  the  size  of  a  box  depends  upon  the  diameter  of  the  cylinder  used.  In 
Mogul,  ten-wheeled,  and  consolidation  engines  the  diameter  of  cylinder  is  also  propor- 
tioned in  accordance  with  the  weight  on  drivers ;  but  under  any  one  of  these  engines 
there  are  more  driving  wheels  than  under  an  eight-wheeled  passenger  engine,  and 
consequently  the  size  of  the  axle  box  suitable  for  an  eight-wheeled  passenger  engine, 
with  cylinders  of  given  diameter,  may  or  may  not  be  suitable  for  a  Mogul,  ten- wheeled, 
or  consolidation  engine  having  cylinders  whose  dimensions  are  equal  to  those  of  the 
cylinders  in  the  passenger  engine.  It  therefore  remains  for  us  to  consider  the  con- 
ditions which  influence  the  size  of  the  journal  in  these  engines,  thereby  enabling 
iis  to  determine  its  dimensions  and  select  the  proper  size  of  driving  box,  from  the 
number  shown  in  Figs.  314  to  340.  In  Mogul,  ten-wheeled,  and  consolidation  engines, 
as  well  as  in  passenger  engines,  the  weight  on  the  journals  must  be  distributed  in 
such  a  manner  as  to  prevent  heating;  and  the  rules  given  for  finding  the  dimen- 
sions of  the  driving  axle  journal  for  a  passenger  engine  may  also,  with  a  slight 
change,  namely,  less  pressure  per  square  inch  on  the  projected  area  of  the  journal,  be 
used  for  computing  the  dimensions  of  driving  axle  journals  in  the  other  classes  of 
engines. 


MODEKX  LOCOMOTirE   COXSTRCCTIOX. 


211 


From  the  foregoing  we  perceive  that  in  all  locomotives  the  driving  axle  journals 
are  proportioned  to  the  pressure  due  to  the  weight  on  the  driving  wheels,  and  when  this 
is  correctly  done,  sufficient  allowance  will  then  have  been  made  for  the  pressure  due 
to  the  load  which  the  engine  has  to  haul ;  we  therefore  leave  the  load  to  be  hauled 
out  of  consideration  in  the  calculations.  In  determining  the  sizes  of  journals  for 
Mogul,  ten-wheeled,  and  consolidation  engines,  we  must  not  lose  sight  of  the  fact  that, 
in  these  classes  of  engines,  particularly  in  Mogul  and  consolidation  locomotives,  the 
driving  wheels  are  smaller  in  diameter  than  those  in  eight-wheeled  passenger  engines, 
and  consequently  the  axles  are  brought  closer  to  the  track  and  exposed  more  to 
the  dust,  thereby  raising  conditions  favorable  for  cutting  the  journals.  Again,  it 
often  happens  that,  on  the  road,  the  counterbalance  weights  are  in  the  way  of  oiling 
the  journals,  and  therefore,  if  the  time  is  limited,  it  may  happen  that  the  oiling  of 
some  of  the  journals  will  be  neglected.  These  classes  of  locomotives  are  also  more  or 
less  looked  upon  as  freight  engines,  and  do  not  always  receive  the  same  care  as 
bestowed  upon  passenger  engines.  For  these  reasons  it  is  always  advisable  to  make 
the  journals  for  freight  engines  comparatively  large  in  diameter.  We  therefore  recom- 
mend that  their  driving  axle  journals  shall  be  so  proportioned  as  not  to  allow  more 
than  160  pounds  per  square  inch  on  the  projected  area  of  the  journal ;  in  fact, 
160  pounds  should  be  the  limit,  instead  of  175  pounds,  as  given  for  passenger  engines, 
and  when  the  design  or  other  given  conditions  will  allow,  the  pressure  per  square 
inch  should  be  even  less  than  160  pounds. 

The  dimensions  of  driving  axle  journals  in  the  following  tables  are  recommended. 
The  pressure  per  square  inch  on  the  projected  area  of  these  journals  will  be  less  than 
160  pounds,  when  the  weight  of  the  engines  correspond  to  the  weights  given  in  Tables 

5,  6,  7,  and  8 : 

TABLE  17. 

DIMENSIONS   OF  DRIVING  AXLE  JOURNALS   FOR  MOGUL  ENGINES. 


Diameter  of  Cylinders. 

Diameter  of  Journal*. 

Length  of  Journals. 

11  inches. 

5  inches. 

6±  inches. 

12      " 

5i 

6f       " 

13       " 

5i 

71      " 

14      " 

6 

7f      " 

15       " 

<H 

7f       " 

16      " 

6f 

8        " 

17       " 

6f 

8         " 

18      " 

7* 

8         " 

19       " 

8 

9        " 

TABLE  18. 
DIMENSIONS   OF  DRIVING   AXLE  JOURNALS   FOR  TEN-WHEELED  LOCOMOTIVES. 


Diameter  of  Cylinders. 

Diameter  of  Journal*. 

Length  of  Journals. 

\-  inches. 

5    inches. 

6J  inches. 

13       " 

5i       " 

6}       " 

14       " 

6         " 

7J      " 

1f>       " 

6 

75.      « 

16      " 

6* 

7J      « 

17       " 

6} 

8 

1H        " 

7i 

8         " 

19       " 

8 

9        " 

212 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


TABLE   19. 
DIMENSIONS   OF  DEIVING  AXLE  JOURNALS   FOR  CONSOLIDATION  LOCOMOTIVES. 


Diameter  of  Cylinders. 

Diameter  of  Journals. 

Length  of  Journals. 

14  inches. 
15       " 
20       " 
24       " 

5    inches. 
5*       " 

74       " 
8         " 

6i  inches. 
7f       " 

8         " 
9        " 

245.  Although  there  are  many  locomotives  with  cylinders  of  the  same  diameters 
as  given  in  these  tables  running  with  smaller  journals,  yet  the  dimensions  given  in 
the  tables  agree  with  the  average  sizes  of  journals  in  modern  locomotives. 

For  the  sake  of  comparison,  we  append  Table  20,  in  which  are  given  the  dimen- 
sions of  a  few  driving  axle  journals  in  modern  locomotives  doing  excellent  service : 


TABLE   20. 


Class  of  Locomotive. 

Dimensions  of  Cylinders. 

Diameter  of  Driving 
Axle  Journal. 

Diameter  of  Main  Driv- 
ing Axle  Journal. 

Length  of  Journals. 

17"  x  24" 

1\  inches. 

7J  inches. 

8   inches. 

18"  x  24" 

8 

8 

gi 

a           it                  n 

18"  x  24" 

g 

8 

12* 

18"  x  24" 

71 

7| 

8 

18"  x  24" 

74 

74 

8 

it           n 

19"  x  24" 

7i 

7$ 

8 

20"  x  24" 

74, 

74 

8 

In  Table  20  it  will  be  seen  that  in  some  of  the  locomotives  the  journals  of  the 
main  driving  axle — that  is,  the  axle  to  which  the  connecting-rods  are  attached — are 
made  \  inch  larger  in  diameter  than  the  other  driving  axle  journals.  Although 
this  is  good  practice,  there  are  master-mechanics  who  object  to  such  an  arrangement, 
because  it  compels  them  to  keep,  for  each  class  of  engine,  two  sizes  of  driving  boxes 
on  hand  ready  to  replace  worn-out  ones,  thereby  increasing  the  stock  which  must 
be  kept  to  facilitate  quick  repairs. 

The  dimensions  of  the  driving  axle  journals  given  in  Tables  17,  18,  and  19  also 
establish  the  sizes  of  driving  boxes ;  and  consequently  we  have  only  to  select,  from  the 
number  of  boxes  shown  in  Figs.  314  to  340,  that  box  which  will  fit  the  journal  to  be 
used. 

When  the  right  driving  box  has  been  selected,  we  then  select  the  correspond- 
ing pedestal  from  the  number  of  pedestals  shown  in  Figs.  271  to  279.  It  should  be 
remembered  that  the  dimensions  of  the  upper  braces  JB2,  and  the  lower  braces  L, 
uniting  the  pedestals,  must  be  computed  to  suit  the  diameters  of  the  cylinders; 
and  therefore  in  frames  for  freight  engines  the  braces  uniting  the  pedestals  will  be 
heavier  than  those  shown  in  the  illustrations.  To  make  this  plain,  we  will  take  an 
example.  Let  it  be  required  to  find  the  principal  dimensions  of  a  frame  for  a 
consolidation  engine  with  cylinders  20  inches  in  diameter.  In  Table  19  we  see  that 
the  driving  axle  journal  must  be  7£  inches  diameter  and  8  inches  long,  hence  the 


MODERN  LOCOMOTITE   CONSTRUCTION.  213 

box  to  be  used  is  that  one  which  is  represented  in  Fig.  317 ;  and  the  pedestal  to  be 
used  for  tliis  box  is  represented  in  Fig.  272.  But  now,  the  dimensions  for  the 
upper  and  lower  frame  braces  must  be  determined  according  to  Rules  33a,  35«, 
and  36,  to  suit  a  cylinder  20  inches  in  diameter,  and  consequently  these  braces  will 
IH-  heavier  than  shown  in  Fig.  272,  but  the  dimensions  of  the  pedestal  itself  will 
remain  as  they  are  given.  There  are  also  a  few  eases  in  which  the  choice  of  the  axle 
box  will  either  compel  us  to  change  the  width  of  frame  suitable  for  a  given  diameter 
of  cylinder  as  established  in  the  illustration  of  pedestals,  or  we  must  increase  the 
width  of  the  axle  box  to  suit  the  given  width  of  frame. 


CHAPTER  VI. 

DRIVING  AXLES.— DRIVING  WHEELS.— COUNTERBALANCE. 

DRIVING  AXLES. 

246.  The  driving  axles  are  made  of  iron  or  steel ;  the  wi'iter  prefers  good  ham- 
mered-iron  axles. 

Some  of  the  different  forms  of  driving  axles,  such  as  are  generally  adopted  under 
all  classes  of  locomotives,  are  shown  in  Figs.  341,  342,  and  343. 

Fig.  341  represents  a  main  driving  axle  suitable  for  eight-wheeled  passenger 
locomotives,  with  cylinders  16  inches  in  diameter ;  a  Mogul  engine  with  cylinders  16 
inches  in  diameter ;  or  a  ten- wheeled  locomotive  with  cylinders  17  inches  in  diameter. 
Those  parts  of  the  axles  marked  A  A  are  generally  called  the  wheel  fits,  and  are 
usually  turned  to  £  inch  less  in  diameter  than  the  journals  of  the  axles  marked 
B  B ;  in  fact,  in  this  style  of  axles  the  difference  between  the  diameter  of  the  wheel 
fit  and  that  of  the  journal  should  never  be  greater  than  J  inch,  as  this  will  give 
a  shoulder  sufficient  for  all  practical  purposes ;  on  the  other  hand,  if  the  difference 
between  these  diameters  is  greater  than  £  inch,  the  axle  will  be  unnecessarily  weak- 
ened, and  will  be  liable  to  break  off  at  the  hub  of  the  wheel.  Sharp  'corners  at  //  are 
another  cause  which  will  lead  to  the  breaking  of  the  axles  near  the  hub,  and  therefore 
sharp  corners  should  not  be  tolerated;  the  junction  between  the  wheel  fit  and  the 
journal  should  always  be  a  curve.  Although  this  manner  of  forming  the  wheel  fit  A, 
that  is,  turning  it  to  a  smaller  diameter  than  that  of  the  journal,  is  quite  a  common 
practice,  we  believe  it  to  be  inferior  to  that  shown  in  Fig.  344.  In  this  design  the 
diameter  of  the  wheel  fit  is  equal  to  that  of  the  journal,  and  consequently  the  strength 
of  the  axle  is  in  nowise  impaired.  The  shoulder  against  which  the  hub  of  the  wheel 
is  pressed  is  formed  in  turning  by  leaving  on  the  axle  a  small  collar  H,  about  -£$ 
inch  larger  in  diameter  than  the  journal,  the  thickness  of  this  collar  being  about 
|  inch  at  the  top.  Here,  also,  sharp  corners  at  i  i  must  be  avoided,  and  the 
junction  between  the  collar  and  axle  nicely  rounded  out.  The  hub  of  the  wheel 
is  counterbored  to  receive  the  collar,  as  shown  in  the  illustration.  This  design  of 
an  axle  has  an  advantage  over  those  shown  in  Figs.  341,  342,  and  343,  namely,  in 
axles  with  wheel  fits  like  that  shown  in  Fig.  344 ;  the  journal  can  be  trued  up  when 
necessary  and  still  leave  the  shoulder  H  unimpaired. 

247.  Sometimes  we  find  the  main  axles  turned  to  equal  diameters  from  hub  to 
hub  of  wheels,  and  frequently  we  find  them  formed  as   shown  in  Fig.  341.     In 


MOVERS  LOCOMOTIVE   COSSTBUCTIOX. 


215 


the  latter  axle  the  central   part  C  is  left  smooth  forged;    the  parts  B  B  are  made 
sufficiently  long  to  receive  the  axle  box  and  eccentrics. 

Fig.  ;>4L'  represents  the  rear  driving  axle  of  an  eight-wheeled  passenger  engine 
with  cylinders  1G  inches  in  diameter.    The  same  size  and  form  of  axle  is  also  used 


*  a 


under  other  engines  under  which  a  main  axle  like  that 
in  Fig.  341  is  used.  In  these  driving  axles  the  central 
part  C  and  the  projections  which  form  the  shoulders 
are  left  smooth  forged;  the  amount  of  projection 
is  equal  to  amount  of  metal  allowed  for  turning  the 
journals  B.  The  cast-iron  collars  G  G  are  either  shrunk 
on  the  axle,  or  held  in  position  by  two  set  screws.  Oc- 
casionally we  find  the  driving  axle  made  as  shown  in 

Fin    'III 

Fig.  343.     The  only  difference  between  the  axle  shown 

in  Fig.  342  and  that  in  Fig.  343  will  be  found  in  the  form  of  the  central  part  CC; 
in  the  latter  the  diameter  is  gradually  reduced  from  the  collar  to  the  center ;  and  in 
the  former  it  is  of  equal  diameter  throughout. 

DRIVING  WHEELS. 

248.  Fig.  345  represents  the  front  view,  Figs.  346,  347,  sections  of  a  driving  wheel, 
and  Figs.  348,  349  represent  sections  of  its  arms.  This  wheel  was  designed  for  and  is 
successfully  used  under  eight-wheeled  fast  passenger  engines  with  cylinders  18  inches 
in  diameter. 

A  driving  wheel  consists  of  two  parts,  namely,  the  driving  wheel  center  marked 
C,  and  the  tire  marked  T.  In  this  country  the  driving  wheel  centers  are  made, 
almost  universally,  of  cast-iron.  Sometimes  the  spokes  are  cast  solid,  but  usually 
they  and  the  rim  are  east  hollow.  The  crank  N  and  the  counterbalance  0  0  form 
part  of  the  wheel  center. 

The  most  common  practice  in  fastening  the  tires  on  the  wheel  center  is  to  shrink 
them  on  the  center.  To  do  this,  the  wheel  centers  are  turned  square  across  (not 


216 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


MI>IH:I;\ 


<  <>\sri;r<"ri<>\. 


217 


tapered),  and  the  tire  bored  out  somewhat  smaller  in  diameter  than  that  of  the  wheel 
•  •enter.  The  tire  is  then  heated,  generally  on  account  of  cleanliness,  by  means  of  a 
number  of  gas  (lames  arranged  for  the  purpose.  When  the  tire,  by  these  means,  has 
l>een  sufficiently  expanded,  it  is  then  slipped  on  the  wheel  center  and  allowed  to  cool, 
thereby  contracting  and  binding  it  firmly  around  the  cast-iron  center. 


TABLE  21. 

STANDARD   SIZES   OF   WHEEL   CENTERS. 


Diameter  of  Wh<,>el  Centers. 

Inside  Diameters  of  Tires. 

38  inches. 
44 
50 
56 
62 
66 

38  inches,  less  0.040 
44       '           "    0.047 
50       '           "    0.053 
56       '           "    0.060 
62       '           "    0.066 
66       '           "    0.070 

A  uniformity  in  the  diameters  of  wheel  centers  has  not  yet  been  thoroughly 
established.  In  1886  the  American  Railway  Master  Mechanics'  Association  recom- 
mended and  adopted  the  above  di- 
ameters as  standards. 

In  this  table  we  see  that,  for 
the  given  diameters  of  wheel  cen- 
ters, the  shrinkage  allowance  in 
the  bore  of  the  tires  is  0.040,  0.047, 
0.053,  0.060,  0.066,  and  0.070  of  an 
inch  respectively,  and  these  are 
claimed  to  be  the  average  of  the 
wide  range  of  shrinkage  allowance 
used  in  actual  practice. 

249.  Figs.  350,  351,  and  352 
represent  the  different  views  of  a 
driving  wheel  designed  for  a  ten- 
wheeled  engine  having  cylinders 
19  inches  in  diameter  ;  Fig.  353  rep- 
resents the  sections  of  the  spokes. 

In  this  wheel  the  spokes  are 
solid  and  the  rim  is  cast  hol- 
low, with  the  exception  of  that 
part  which  forms  the  counter- 
balance 000,  extending  from  u 
(Fig.  250)  to  an  equal  distance  on 
the  other  side  of  the  center  line; 


this  part  of  the  rim  is  cast  solid. 


Fig.  353     Fig.  351  Fig.  352 


In  order  to  avoid  a  shrinkage  stress,  the  counterbalance  is  parted  at  s  s;  these  open- 
ings extend  to  the  rim,  but  not   through  it;    the  latter  is  parted  or  cored  through 


218  MODERN  LOCOMOTIVE   CONSTRUCTION. 

at  r  and  in  a  corresponding  place  on  the  other  side  of  the  center  line.  The  openings 
r  are  generally  slotted  and  cast-iron  liners  driven  in. 

In  this  wheel  the  whole  crank  is  cored  out,  the  cored  part  extending  to  the  axle, 
and  leaving  all  around  it  an  opening  1|  inches  wide.  There  is  an  objection  to  this 
opening :  it  will  interfere  with  the  guidance  of  the  axle  when  it  is  to  be  pressed  into 
the  hub  of  the  wheel.  It  seems  to  us  that  a  few  ribs  cast  into  the  core  opening  around 
the  axle  will  be  an  improvement,  by  which  considerable  annoyance  may  sometimes  be 
avoided. 

It  will  be  noticed  that  the  hub  of  the  wheel  is  counterbored  at  b ;  the  reason  for 
doing  so  is  to  allow  the  full  diameter  of  the  axle  to  extend  into  the  wheel,  bringing  the 
shoulder  of  the  wheel  fit  inside  of  the  hub,  instead  of  against  the  hub,  as  explained  in 
Art.  247.  The  object  of  this  design  is  to  prevent  the  breaking  of  the  axle,  which 
occasionally  occurs  when  the  shoulder  of  the  wheel  fit  is  pressed  against  the  outside  of 
the  hub. 

The  ribs  m  m,  shown  in  Figs.  345  and  350,  are  for  the  purpose  of  stiffening  the 
hollow  rims. 

In  some  wheels  the  ribs  are  placed  at  the  end  of  each  spoke,  as  shown  in  Fig.  350 ; 
in  other  wheels  they  are  placed  at  the  end  of  each  spoke  and  midway  between  them, 
as  shown  in  Fig.  345. 

250.  Figs.  354  to  357  inclusive  show  different   views   of  a  driving  wheel  de- 
signed for  an  eight-wheeled  passenger  engine  having  cylinders  19  inches  in  diameter. 
In  this  wheel  the  spokes  and  rim  are  cast  solid.     One  peculiarity  of  this  wheel,  not 
often  found  in  others,  is  that  the  rim  of  the  wheel  center  has  a  shoulder  against 
which  the  tire  is  pressed.     The  object  of  this  shoulder  is  to  prevent  the  tire  from 
slipping  inwards  when  the  flange  is  working  against  the  rail.     At  first  sight,  it  may 
appear  that  a  shoulder  of  this  kind  is  unnecessary,  but  when  locomotives  are  fitted 
up  with  driver  brakes,  and  these   applied,  the  tire  will  in   some  instances  become 
sufficiently  hot  to  expand  and  thereby  loosen  it,  and  hence  the  importance  of  the 
shoulder  will  be  apparent. 

We  also  notice  that  for  these  wheels  the  wheel  fit  is  tapered ;  its  large  diameter  is 
the  same  as  that  of  the  journal,  and  consequently  there  are  no  shoulders,  such  as 
shown  in  Figs.  341  and  342.  This  is  another  method  sometimes  adopted  for  the 
prevention  of  breaking  the  axles  near  or  at  the  hub  of  the  wheel. 

The  rim  of  this  wheel  is  cored  through  in  two  places  r  r,  and  at  the  center  rz  of 
the  counterbalance.  In  our  opinion,  the  positions  of  the  openings  r  r  r2  are  better 
located  than  the  two  openings  shown  in  Fig.  350 ;  because  in  Fig.  354  the  distances 
between  the  openings  are  nearly  equal,  and  the  openings  r  r  through  the  rim  are 
placed  between  such  spokes  where  the  strength  of  the  hub  is  reinforced  by  ribs  cast 
between  the  spokes,  the  result  being  that  in  this  wheel  the  openings  through  the  rim 
will  not  widen  as  easily  nor  as  much  as  the  openings  in  the  wheel  shown  in  Fig.  350, 
when  the  axle  is  forced  into  the  hub  of  the  wheel 

251.  Figs.  358  to  362  inclusive  represent  a  driving  wheel  designed  for  an  eight- 
wheeled  passenger  engine  having  cylinders  18  inches  in  diameter.     This  wheel  has 
solid  spokes.     The  rim  is  cast  hollow,  but  differs  from  those  previously  shown,  in  the 
fact  that  in  this  wheel  the  cored  part  in  the  rim  does  not  extend  to  the  periphery ;  it 


Mnlii:i;\  LOCOMOTIf'E   CONSTRUCTION. 


219 


220 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


is  closed,  as  shown  in  Fig.  362,  and  indicated  by  the  dotted  lines  in  Fig.  358.  The 
only  openings  in  the  rim  of  this  wheel  are  the  core  holes  w  w ;  ribs  in  the  rim,  similar 
to  those  marked  m  m  in  Figs.  345  and  350,  are  not  used.  The  rim  from  u  to  u  is  cast 
solid,  but  the  extra  metal  thus  obtained  is  not  sufficient  for  counterbalancing,  and 
therefore  two  or  four  separate  pairs  of  counterbalance  weights,  as  the  case  may 


Section  of  Spokr  through  A. 

"T 


Section  of  Spoke  through  Ji. 


r_ J 

Section  of  Hint  through  C. 


<r  —  -ax  ---  4-;  —  2*  ----  H 


JL- 


-  53%  betwten  fire* 


Fig.  389 


require,  are  bolted  between  the  spokes.  These  weights,  and  the  manner  of  bolting 
them  to  the  wheel,  are  illustrated  in  Figs.  388  and  389.  The  hub  of  the  wheel  is 
cored  out  at  a  a,  having  small  core  openings  v  v  extending  to  the  axle. 

In  this  wheel  the  spokes,  as  they  extend  towards  the  hub,  incline  outwards; 
wheels  of  this  kind  are  sometimes  called  dished  wheels.  The  object  of  placing 
the  spokes  in  this  position  is  to  make  room  for  a  comparatively  wide  axle  box; 
but  there  is  a  limit  to  the  inclination  of  the  spokes,  for  if  they  are  brought  out  too 
far,  it  will  be  necessary  to  spread  the  cylinders  also,  which  is  always  an  objectionable 
feature. 

252.  The  wheels  are  usually  forced  on  the  axle  with  a  hydraulic  press.  The 
pressure  should  be  equal  to  about  9  tons  per  inch  of  diameter,  so  that  an  axle  6  inches 
in  diameter  will  be  forced  into  the  wheel  with  a  pressure  of  6  x  9  =  54  tons ;  or  an 
axle  7  inches  in  diameter,  with  a  pressure  of  7  x  9  =  63  tons. 

To  add  further  security,  keys  are  driven  into  the  wheel  and  axle.  These  keys  are 
generally  made  J  inch  square  for  axles  less  than  6  inches  in  diameter ;  1  inch  square 


MODES*  LOCOMOTIVE   CONSTRUCTION.  221 

for  axles  varying  from  6  to  7  inches  in  diameter ;  and  14  inches  square  for  axles 
7  indies  ami  over  in  diameter. 

The  illustrations  represent  driving  wheels  such  as  are  used  on  some  of  our  promi- 
nent railroads;  they  have  been  taken  directly  from  working  drawings  kindly  given 
to  the  writer  by  locomotive  builders  for  the  purpose  of  illustrating  these  pages.  The 
designs  of  these  wheels  also  indicate  the  study  and  care  bestowed  upon  this  subject  by 
master-mechanics,  so  as  to  obtain  a  strong  and  safe  wheel,  thereby  greatly  promoting 
the  safety  of  the  passengers: 

•-!.">:!.  Figs.  363  and  364  represent  the  two  views  of  a  driving  wheel  designed  for  a 
consolidation  engine  having  cylinders  20  inches  in  diameter.  Fig.  365  represents  a 
section  through  the  rim,  and  Figs.  366,  367  represent  sections  of  one  of  the  arms. 
These  are  cast  solid,  the  rim  hollow,  its  form  being  similar  to  that  of  the  rim  shown  in 
Fig.  358.  The  counterbalance  weight  is  cast  solid  throughout,  without  any  openings 
to  prevent  shrinkage  stress. 

Figs.  368  and  369  represent  two  views  of  another  driving  wheel  designed  for  a 
consolidation  engine  with  cylinders  20  inches  in  diameter.  The  construction  of  this 
wheel  is  somewhat  different  from  that  shown  in  Fig.  363 ;  it  has  hollow  spokes  and  a 
hollow  rim,  the  hollow  part  of  the  rim  extending  to  the  periphery  of  the  wheel  and 
strengthened  by  the  ribs  m  m;  the  counterbalance  has — for  the  purpose  of  avoiding 
shrinkage  stress — openings  r  r  cast  into  it,  and  extending  to  the  rim. 

In  Fig.  368  we  see  that  the  inner  face  of  the  wheel  center  projects  beyond  the 
inner  face  of  the  tire. 

This  style  of  wheel  is  adopted  for  locomotives  designed  for  roads  having  tracks 
of  5  feet  gauge,  which  in  the  near  future  are  to  be  reduced  to  a  gauge  of  4  feet  8£ 
inches.  Now,  locomotives  designed  for  a  5  feet  gauge,  and  having  the  tires  placed  on 
the  wheel  centers,  as  shown  in  Fig.  368,  can  readily  be  changed  to  suit  a  gauge  of  4  feet 
8£  inches,  as  all  that  need  be  done  is  to  heat  the  tires  and  move  them  inwards  on 
the  wheel  centers,  and  then  turn  off  the  outer  face  of  the  latter  to  suit  the  tires. 

COUNTERBALANCE. 

254.  In  traveling  in  a  railroad  car  it  may  sometimes  be  noticed  that  the  motion  of 
the  train  is  not  uniform,  but  is  accompanied  by  jerks  occurring  at  regular  intervals. 
This  kind  of  irregular  motion  is  often  due  to  the  locomotive,  which  is  imperfectly 
counterbalanced.  Consequently,  we  may  say — in  a  general  way — that  in  a  locomotive, 
all  parts  whose  weights  have  a  bad  influence  on  the  smooth  forward  and  backward 
motion  of  the  engine,  and  tend  to  produce  jerks  in  its  motion,  must  be  counter- 
balanced; and  therefore  not  only  the  weight  of  the  crank  and  its  pin,  and  other  parts 
attached  to  the  crank-pin  which  have  a  rotary  motion,  must  be  counterbalanced,  but 
also  those  parts  which  are  attached  to  the  crank-pin  and  haves  a  reciprocating  motion. 
Hence  in  locomotives  the  weights  of  the  cranks,  pins,  connecting-  and  side-rods, 
piston-rods,  crossheads,  and  pistons  must  be  counterbalanced.* 

In  order  to  explain  as  clearly  as  possible  the  method  of  counterbalancing  these 

•  A  difference  of  opinion  exists  in  regard  to  the  amount  of  weight  to  be  counterbalanced,  as  will  be  seen 
farther  on. 


222 


MODERN  LOCOMOTIVE   CONSTBVCT1OX. 


UODERX  LOCOMOTirE   CONSTRUCTION. 


223 


axle 


224  MODERN  LOCOMOTIVE   CONSTRUCTION. 

parts  of  the  engine,  and  reduce  this  subject  to  simple  problems,  we  will  first  consider 
the  principles  upon  which  the  method  is  based,  and  the  application  of  these  principles 
to  the  method  used  for  finding  the  counterbalance  for  objects  of  simple  outline. 

In  the  first  place,  then,  we  will  determine  the  amount  of  counterbalance  required 
for  a  crank  such  as  is  shown  in  Fig.  370.  This  crank  is  supposed  to  be  of  equal 

thickness  throughout,  and  its  form  is  perfectly  sym- 
metrical— that  is  to  say,  the  diameter  of  the  hole  for 
the  axle  is  equal  to  that  of  the  hole  for  the  crank- 
pin,  the  ends  of  the  crank  are  exactly  alike,  and  the 
depth  c  d  or  e  /the  same  throughout. 

ria  370  Assume  now  that   the  crank   stands  in  a  hori- 

zontal position,  as  shown  in  our  illustration,  and  that 

it  is  divided  by  vertical  planes  represented  by  the  lines  c  d,  ef,  etc.,  into  any  number 
of  parts ;  then  the  weight  of  each  part  with  a  leverage  corresponding  to  the  distance 
between  it  and  the  center  a  of  the  axle  will  act  with  a  certain  amount  of  energy, 
tending  to  turn  the  axle  around  its  center,  bringing  the  center  line  a  b  of  the  crank 
in  a  vertical  position  below  the  axle. 

Now,  in  order  to  enable  us  to  compare  readily  the  amount  of  this  energy  with  that 
of  a  force  applied  to  some  other  point  at  a  given  distance  from  the  center  a,  we  assume 
that  the  weights  of  the  different  parts  c  d,  ef,  etc.,  of  the  crank  are  concentrated  on 
the  line  a  b  at  a  single  point  i,  so  that  the  same  amount  of  energy  be  developed  as  is 
developed  by  distributing  the  weight  of  the  parts  along  the  center  line  a  b ;  this  point 
i  will  coincide  with  the  center  of  gravity  of  the  crank.  Hence  our  first  step  will  be  to 
determine  this  center  of  gravity. 

The  center  of  gravity  of  every  solid  or  body  is  a  point  about  which  all  the  parts 
of  the  solid  acted  upon  by  the  force  of  gravity  balance  each  other,  so  that,  if  the  solid 
be  suspended  from  that  point  (center  of  gravity),  the  solid  will  be  in  equilibrium  in 
any  position  it  may  be  placed. 

Since  the  crank  represented  in  Fig.  370  is  symmetrical,  its  center  of  gravity  i 
must  lie  in  the  center  line  a  ft,  and  midway  between  the  centers  a  and  I.  Hence  in 
this  case  the  center  of  gravity  i  is  obtained  without  any  calculation. 

After  the  center  of  gravity  has  been  determined,  we  may  assume  that  the  distance 
between  the  center  a  of  the  axle  and  the  center  of  gravity  i  represents  the  length  of 
an  arm  of  a  lever  whose  fulcrum  is  at  the  center  a.  If  the  whole  weight  of  the  crank 
is  applied  to  the  extremity  i  of  the  lever  arm  a  i,  then  the  effect  produced  or  the 
energy  due  to  the  weight  applied  to  the  end  of  the  lever  arm  will  be  equal  to  the 
energy  of  the  distributed  weight  of  the  crank.  It  is  the  influence  of  this  energy,  in 
counterbalancing  the  weight  of  the  crank,  that  is  to  be  destroyed  by  a  force  whose 
energy  is  equal  and  acting  opposite  to  that  due  to  the  weight  of  the  crank. 

By  the  assumption  that  the  whole  weight  of  the  crank  is  concentrated  at  one 
point — the  center  of  gravity — or  applied  to  a  point  coinciding  with  the  center  of 
gravity,  we  obtain  an  easy  way  of  comparing  the  energy  developed  by  the  weight  of 
the  crank  and  that  developed  by  the  counterbalance,  and  also  an  easy  way  for  deter- 
mining the  number  of  pounds  of  metal  required  in  the  latter. 

In  determining  the  counterbalance  we  must  also  find  its  center  of  gravity,  and,  as 


MODKRX  LOCOMOTIVE   CONSTRUCTION.  225 

in  the  case  of    the  crank,  assume  its  weight  to  be  concentrated  at  its  center  of 
gravity. 

Those  conditions  will  reduce  the  whole  method  of  counterbalancing  to  a  simple 
problem,  in  which  the  given  conditions  are  such  as  represented  in  Fig.  371.  In  this 
figure  the  point  a  represents  the  center  of  the  axle ;  the  line  k  b,  passing  through  the 
center  rt,  represents  a  lever  whose  fulcrum  is  at  a ;  the  line  a  b  is  the  arm  of  the  force  C, 
its  length  is  equal  to  the  distance  between  the 
center  of  the  axle  and  the  center  of  gravity 
of  the  crank.  The  line  a  k  is  the  arm  of  the 


force  R,  its  length  is  equal  to  the  distance  be-     '   "    ' 

tween  the  center  of  the  axle  and  the  center 

of  gravity  of  the  counterbalance.     The  weight  of  the  crank  is  represented  by  C  which 

is  attached  to  the  point  ft;  the  weight  of  the  counterbalance  is  represented  by  .R, 

which  is  attached  to  the  point  k. 

Now,  since  in  all  levers  of  this  kind  which  are  in  equilibrium  the  product  ob- 
tained by  multiplying  the  length  of  the  arm  a  b  by  the  weight  6',  which  is  suspended 
from  this  arm,  must  be  equal  to  the  product  obtained  by  multiplying  the  length  of  the 
arm  a  k  by  the  weight  It  suspended  from  it,  we  can  find  the  number  of  pounds  of 
inetal  required  in  the  counterbalance  by  the  following  rule : 

RULE  42. — Multiply  the  distance  between  the  center  of  the  axle  and  the  center  of 
gravity  of  the  crank  in  inches  by  the  weight  of  the  crank  in  pounds,  and  divide  this 
product  by  the  distance  between  the  center  of  the  axle  and  the  center  of  gravity  of 
the  counterbalance  in  inches ;  the  quotient  will  be  the  weight  of  the  counterbalance  in 
pounds. 

EXAMPLE  74. — The  length  of  the  crank  (Fig.  370)  from  the  center  a  to  the  center 
b  is  12  inches ;  the  weight  of  the  crank  is  300  pounds.  It  is  required  to  find  the  weight 
of  the  counterbalance,  which  is  placed  so  that  the  distance  between  its  center  of 
gravity  and  the  center  of  the  axle  is  9  inches. 

Since  the  form  of  this  crank  is  symmetrical,  and  since  its  length  is  12  inches,  the 
distance  between  its  center  of  gravity  and  the  center  of  axle  is  6  inches ;  hence  we 
have, 

6  x  300 

,.          =  200  pounds; 

which  is  the  weight  necessary  for  counterbalancing  the  weight  of  the  crank  only. 

255.  In  order  to  express  concisely  that  which  is  to  follow,  it  will  be  necessary  to 
give  a  more  general  definition  of  the  term  "  arm  of  the  force,"  or  simply  "  arm,"  and 
also  a  definition  of  the  term  "  moment  of  a  force  "  which  we  shall  introduce. 

Fig.  371.  Let  k  b  represent  a  lever,  a  its  fulcrum,  It  and  C  weights  attached  to 
the  ends  k  and  b  of  the  lever. 

By  the  term  "  arm  of  the  force  "  is  meant  the  perpendicular  distance  from  the  ful- 
crum to  the  line  of  direction  of  a  force  applied  to  the  arm.  For  instance:  The  weight 
C  when  applied  to  the  end  I  of  the  lever  will  act  in  a  vertical  line  b  C,  and  therefore, 
according  to  our  definition,  the  line  a  6,  which  is  perpendicular  to  the  line  I  C,  will 
If  the  arm  of  the  force  C.  The  line  a  k  is  the  arm  of  the  weight  R,  because  <t  /,-  is  the 
perpendicular  distance  from  the  fulcrum  a  to  the  line  k  R  in  which  the  weight  It 


226  MODERN  LOCOMOTIVE   CONSTRUCTION. 

acts.  In  fact,  we  may  say  that  the  arm  of  the  force  is  the  shortest  line  that  can  be 
drawn  from  the  fulcrum  to  the  line  of  direction  of  the  force. 

By  the  term  "  line  of  direction  of  a  force  "  is  meant  a  line  indicating  the  direction 
in  which  the  force  acts  ;  the  length  of  this  line  is  not  limited  by  the  distance  between 
the  arm  and  the  weight  attached  to  it.  Thus,  in  Fig.  372,  let  the  lever  on  one  side 
of  the  fulcrum  be  bent  as  indicated  by  the  line  d  a.  In  this  case  the  line  of  direction 
in  which  the  force,  due  to  the  weight  R,  acts  is  not  limited  by  the  end  of  the  lever 
d  and  the  weight  R,  but  the  line  of  direction  is  represented  by  the  line  d  e  extend- 
ing below  the  fulcnim,  so  that  the  line  a  k  can  be  drawn  from  the  fulcrum  a  per- 
pendicular to  the  line  of  direction  d  e.  In  this  case  the  line  a  &  is  the  ami  of  the 
weight  R. 

Since  the  weight  E  acts  with  a  leverage  a  k,  we  say  that  a  k  is  IPs  arm,  and  for  a 
similar  reason  we  say  that  a  b  is  C's  arm. 

MOMENT   OF  A  FORCE. 

256.  The  moment  of  a  force  with  respect  to  a  point  is  the  product  obtained  by 
multiplying  the  force  by  the  perpendicular  distance  from  the  point  to  the  line  of 
direction  of  the  force.  When  the  forces  are  applied  to  a  lever,  then  the  product  of 
each  force  multiplied  by  its  arm  is  the  moment  of  that  force.  Thus  : 

In  the  lever  k  b,  Fig.  371,  the  force  due  to  the  weight  R  tends  to  turn  the  lever 
around  the  point  or  fulcrum  a,  and  the  same  thing  may  be  said  of  the  force  due  to  C, 
that  is,  it  tends  to  turn  the  lever  around  the  same  fulcrum  «,  but  in  an  opposite 
direction.  The  weight  E  acts  with  a  leverage  a  £,  hence  the  product  of  the  weight  R 
multiplied  by  its  arm  a  k  is  the  moment  of  the  force  due  to  the  weight  R.  In  like 
manner,  the  product  obtained  by  multiplying  the  weight  C  by  its  arm 
a  b  is  the  moment  of  the  force  C. 

The  moment  of  a  force  is  used  as  a  measure  of  its  tendency  to 
turn   the    lever   around   a  point,   or  the    fulcrum  a.     By  establish- 
ing this  measure  we  obtain  an  easy  method  for 
comparing  the  effect  of  the  two  forces  applied  to 


Fig  372  fc^]      a  ^ever  5  ^a^  *s  *°  sav>  we  can  1>ea(lily  determine 

whether  the  forces  applied  to  a  lever  will  hold  it  in 
equilibrium  or  not  ;  and  if  these  forces  do  not  hold  the  lever  in  equilibrium,  we  are 
enabled  to  calculate  quickly  the  amount  by  which  one  of  the  forces  must  be  increased 
or  decreased,  and  this  is  exactly  what  we  have  to  do  in  counterbalancing  some  of  the 
weights  in  a  locomotive. 

When  a  lever  is  in  equilibrium,  the  moments  of  the  forces  are  equal  to  each  other. 
To  make  the  foregoing  principles  clear  let  us  take  the  following  example  : 
EXAMPLE  75.  —  Suppose  that  in  Fig.  371  the  length  of  the  arm  a  k  is  two  feet,  and 
the  weight  R  attached  to  it  is  150  pounds,  the  length  of  the  lever  arm  a  b  is  four  feet, 
and  the  weight  C  is  75  pounds.     Will  the  lever  k  b  under  these  conditions  be  in  equi- 
librium ?     The  moment  of  the  force  due  to  R  is  equal  to  the  product  of  the  weight  R 
into  its  arm  a  k;  hence  we  have:  150  pounds  x  2  feet  =  300  foot  pounds  =  moment 
of  R. 


\  LOCOMOTIVE   CONSTRUCTION.  227 

The  moment  of  the  force  C  is  equal  to  the  pi-oduct  of  the  weight  C  into  its  arm 
a  //,  hence  \ve  have:  75  pounds  x  4  feet  =  300  foot  pounds  =  moment  of  C. 

Here,  then,  we  see  that  the  moments  of  the  weights  R  and  C  are  equal,  and  conse- 
quently the  lever  must  be  in  equilibrium.  Let  us  take  another  example. 

EXAMPLE  76. — The  length  of  the  arm  a  k  in  Fig.  371  is  two  feet,  the  weight  R 
attached  to  it  is  300  pounds ;  the  length  of  the  arm  a  b  is  four  feet,  and  the  weight 
('  attached  to  it  is  75  pounds.  Will  the  lever  under  these  conditions  be  in  equi- 
librium? If  not,  what  change  must  be  made  in  the  weight  R1 

The  moment  of  the  weight  R  is  equal  to  300  pounds  x  2  feet  =  600  foot 
pounds. 

The  moment  of  the  weight  C  is  equal  to  75  pounds  x  4  feet  =  300  foot 
pounds. 

Here  we  see  that  the  moment  of  R  is  600,  and  the  moment  of  C  is  300,  and  since 
the  moments  are  not  equal,  the  lever  cannot  be  in  equilibrium.  This  also  indicates, 
that  because  the  moment  of  R  is  greater  than  the  moment  of  C,  the  weight  of  R  is  too 
great,  and  consequently  it  will  turn  the  lever  around  the  point  or  fulcrum  a  and  pull 
the  end  k  downwards. 

In  order  to  produce  equilibrium,  we  would  have  to  change  one  of  the  arms,  or 
change  one  of  the  weights;  but  according  to  the  conditions  given  in  our  example, 
we  can  make  only  one  change,  and  that  is  in  the  weight  R.  Hence,  our  next  step  will 
be  to  determine  by  calculation  the  amount  of  reduction  in  the  weight  R. 

We  have  seen  that  in  order  to  produce  equilibrium  the  moments  of  the  two 
forces  must  be  equal.  We  know  that  the  moment  of  the  weight  C  is  300,  and  we  also 
know  that  the  length  of  Rs  arm  is  2  feet;  now,  since  the  product  of  2  feet  into 
the  weight  R,  that  is,  the  moment  of  72,  must  be  equal  to  300  to  produce  equi- 
librium, we  simply  divide  the  moment  of  C  by  the  length  of  R?s  arm  and  obtain 

300 
-  =  150  pounds  for  the  weight  of  R.    Here  we  see  that  R  must  be  reduced  to  one- 

m 

half  of  its  original  weight. 

In  calculating  the  moments  we  must  always  use  the  same  unit  of  length  for  both 
lever  arms,  and  also  the  same  unit  of  weight  for  the  forces  applied  to  the  lever.  That 
is  to  say,  when  we  multiply  the  arm  a  k  in  feet  by  the  weight  of  R  in  pounds,  we 
must  also  multiply  the  arm  a  b  in  feet  (not  inches)  by  the  weight  of  C  in  pounds. 
If  we  multiply  the  arm  a  k  in  inches  (which  we  are  at  perfect  liberty  to  do)  by 
the  weight  R  in  pounds,  then  we  must  also  multiply  the  lever  arm  a  b  in  inches  (not 
feet)  by  the  weight  C  in  pounds.  Or,  if  we  adopt  ounces  as  the  unit  of  measurement 
for  the  weight  R,  we  must  also  adopt  ounces  for  the  unit  of  measurement  for  the 
weight  C.  To  make  this  plain,  let  us  consider  the  conditions  given  in  Example  76, 
namely,  that  the  arm  a  k  is  equal  to  2  feet,  the  arm  a  b  equal  to  4  feet,  the  weight 
R  equal  to  300  pounds,  and  the  weight  C  equal  to  75  pounds.  Taking  feet  as  the 
unit  of  measurement  for  the  length  of  the  arms,  we  have,  for  the  moments  of  R  and  C: 

300  x  2  =  600  foot  pounds  =  moment  of  R ;  and 
75  x  4  =  300  foot  pounds  =  moment  of  C. 

Here  we  see  that  the  moment  of  R  is  equal  to  twice  that  of  C. 


228  MODERN  LOCOMOTIVE   CONSTRUCTION. 

Taking  inches  as  the  unit  of  measurement  for  the  length  of  the  arms,  we  have 
for  the  moments  of  E  and  C: 

300  x  24  =  7200  inch  pounds  =  moment  of  E ;  and 
75  x  48  =  3600  inch  pounds  =  moment  of  C. 

Here,  again,  the  moment  of  E  is  equal  to  twice  that  of  C.  Hence  we  see  that,  for  the 
purpose  of  comparing  the  effects  of  the  forces  applied  to  the  lever,  it  makes  no  differ- 
ence whether  we  adopt  feet  or  inches  for  the  unit  of  measurement,  so  long  as  we  keep 
the  same  unit  for  both  arms. 

If  these  principles  are  understood,  considerable  of  the  difficulty  in  determining  the 
counterbalance  for  an  engine  will  disappear. 

AMOUNT  OF  COUNTERBALANCE. 

257.  In  Art.  254  it  was  seen  that  in  determining  the  counterbalance  for  the  crank, 
we  simply  assumed  the  line  drawn  from  the  center  of  gravity  of  the  crank  to  the  cen- 
ter of  gravity  of  the  counterbalance  to  represent  a  lever  with  the  fulcrum  at  the  center 
of  the  axle,  and  then  calculated  the  weight  attached  to  one  end  of  the  lever  that  would 
counterbalance  the  total  weight  of  the  crank  attached  to  the  opposite  end  of  the  lever. 
As  we  proceed,  it  will  be  seen  that,  for  the  sake  of  convenience  in  calculating 
the  total  amount  of  counterbalance  required  in  a  locomotive,  it  is  desirable  to  adopt,  in 
place  of  the  whole  weight  of  the  crank  applied  to  its  center  of  gravity,  a  smaller 
weight  applied  to  the  center  of  the  crank-pin,  which  will  have  the  same  effect  as  the 
whole  weight  of  the  crank  applied  to  its  center  of  gravity.  This  smaller  weight  is 
determined  in  the  following  manner : 

EXAMPLE  77. — Fig.  373.  Let  a  represent  the  center  of  the  axle,  and  let  the  horizontal 
line  k  d  drawn  through  a  represent  a  lever  with  its  fulcrum  at  a ;  also  let  a  d  repre- 
sent the  length  of  the  crank — that  is, 
the  distance  between  the  center  a  of 
the  axle  and  the  center  d  of  the  crank- 
||      pin;  b  the  center  of  gravity  of  the 
crank;    C  the   whole  weight  of  the 
crank  applied  to  b ;  and  E  the  weight 
of  the  counterbalance  applied  to  the 
point  k  coinciding  with  the  center  of 
countenance  -—  •"•»»  gravity  of  the  counterbalance.    Let 

the  total  weight  of  the  crank  be  300 
jf'iff,  o  /  *> 

pounds,  the  length  of   the  crank  12 

inches,  and  the  distance  from  the  center  of  axle  «  to  the  center  of  gravity  k  of  the 
counterbalance  9  inches,  and  the  distance  from  the  center  of  axle  to  the  center  of 
gravity  of  the  crank  6  inches;  it  is  required  to  determine  by  calculation  the  weight  W 
applied  to  the  center  d  of  the  crank-pin ;  the  weight  W  is  to  have  the  same  effect 
or  tendency  to  turn  the  crank  around  the  center  a  of  the  axle  as  that  of  the  whole 
weight  of  the  crank  (300  pounds)  applied  to  the  center  of  gravity  b. 

We  have  seen  in  Example  74  (Art.  254)  that  to  counterbalance  the  weight  of  this 


MODERX    LOCOMOTIVE    COySTKUCTIOX.  229 

crank  under  the  given  conditions,  we  require  200  pounds,  and  this  countei'balance 
cannot  be  changed  as  long  as  the  weight  of  the  crank,  its  center  of  gravity,  and  the 
center  of  gravity  of  the  counterbalance  remain  as  they  are  given. 

But  we  have  already  seen  that  the  moment  of  a  force  is  a  measure  of  its  tendency 
to  produce  rotation.  The  moment  of  the  counterbalance  R  in  our  example  is  equal  to 
I'll!)  x  f)  =  1800  inch  pounds;  and  if  the  lever  A:  b  is  to  remain  in  equilibrium,  the 
moment  of  weight  6'  must  also  be  equal  to  1,800  inch  pounds,  which  we  find  to  be  the 
case  in  this  example,  for  the  distance  from  the  center  a  of  the  axle  to  the  center  of 
gravity  of  the  crank  is  G  inches,  and  the  weight  of  the  crank  is  300  pounds,  hence  the 
moment  of  C  is  300  x  0  =  1800  inch  pounds. 

If  we  now  replace  the  weight  C  of  the  crank  applied  to  its  center  of  gravity 
by  another  weight  W  applied  to  the  center  of  the  crank-pin  d,  and  at  the  same  time 
preserve  an  equilibrium,  it  will  easily  be  perceived  that  the  moment  of  W  must 
be  equal  to  the  moment  of  R.  Again,  since  the  moment  of  C,  as  we  have  shown,  is 
equal  to  the  moment  of  R,  we  may  say  that  the  moment  of  W  must  be  equal  to  the 
moment  of  C.  Thei'efore,  when  to  the  center  d  of  the  crank-pin  a  weight  W  is  to  be 
attached,  which  shall  have  the  same  tendency  to  turn  the  crank  around  the  center  a  of 
the  axle,  we  have,  for  determining  the  amount  of  the  weight  IF,  the  following  nile : 

KULE  43. — Multiply  the  distance  from  the  center  of  gravity  of  the  crank  to  the 
center  of  the  axle  in  inches  by  the  total  weight  of  the  crank  in  pounds ;  divide  this 
product  by  the  length  of  the  crank  in  inches;  the  quotient  will  be  the  number  of 
pounds  in  the  weight  W. 

Therefore,  the  required  weight  W  in  our  example  will  be 

6  inches  x  300  pounds 

— TTT; — c~  ~  =  150  pounds 

12  inches 

Generally  in  locomotives  the  center  of  gravity  of  the  crank  is  not  in  the  center  of 
its  length ;  but  Eule  43  will  apply  to  all  cases  in  which  the  center  of  gravity  is  in  the 
center  of  the  length  or  otherwise.  How  the  center  of  gravity  in  bodies  of  different 
forms  can  be  obtained  will  soon  be  explained. 

The  advantage  of  determining  the  weight  W  applied  to  the  center  of  crank-pin  in 
place  of  the  whole  weight  C  of  the  crank  applied  to  its  center  of  gravity,  is,  that  the 
counterbalance  for  the  crank  and  other  parts  attached  to  the  crank-pin,  which  must 
also  be  counterbalanced,  can  be  found  with  less  labor. 

POSITION   OF  THE  CENTER  OF  GRAVITY  OF  THE  COUNTERBALANCE. 

258.  So  far  we  have  assumed  that  the  center  of  gravity  of  the  crank  and  that 
of  the  counterbalance  lie  in  one  straight  line  passing  through  the  center  of  axle.  We 
have  also  stated  that  the  length  of  the  arm  of  the  counterbalance  is  equal  to  the  dis- 
tance between  the  center  of  the  axle  and  the  center  of  gravity  of  the  counterbalance. 
It  is  now  to  be  explained  how  the  position  of  the  center  of  gravity  of  the  counterbal- 
ance can  be  determined. 

Let  the  full  lines  in  Figs.  374  and  375  represent  two  views  of  a  counterbalance; 
it  is  required  to  find  the  distance  between  its  center  of  gravity  and  center  of  the 


230 


MODERN    LOCOMOTIVE    CONSTRUCTION. 


axle ;  the  arc  /  b  tj  is  described  from  the  center  of  the  axle,  and  its  radius  is  11 
inches.  The  middle  point  b  of  the  arc  fb  g,  the  center  of  gravity  of  the  weight,  and 
the  center  of  axle  are  to  lie  in  one  straight  line. 

From  the  conditions  given,  we  know  that  the  distance  from  the  point  I  to  the 
center  of  axle  is  11  inches ;  it  therefore  remains  to  find  only  the  distance  between  I  and 
the  center  of  gravity  of  the  weight. 

Fig.  375  shows  that  the  thickness  of  the  weight  is  the  same  throughout ;  we  have, 
therefore,  two  methods  for  finding  the  position  of  its  center  of  gravity :  first,  a  geomet- 
rical method ;  and  second,  a  practical  method. 


GEOMETEICAL   METHOD. 


For  the  purpose  of  finding  the  distance  between  the  center  of  gravity  and  the 
point  b  when  the  thickness  of  the  weight  is  uniform,  we  have  only  to  find  the  center  of 
gravity  of  the  surface  or  plane  e  a  d g  If,  Fig.  374. 

Fig.  374.  Join  the  points  e  and  d  by  a  straight  line,  and  bisect  this  line  by  the 
perpendicular  line  a  b,  cutting  the  arc  /  g  in  the  point  I.  The  line  a  b  will  divide 


ill  tnclies  tu 
centre  of  axle 


Fig.  375 


fig.  374 


the  plane  e  a  d  g  b  f  into  two  equal  parts,  and  is  therefore  a  center  line ;  it  also 
contains  the  center  of  gravity  of  the  plane,  because  it  passes  through  the  centers  of 
all  lines  drawn  parallel  to  the  line  e  d.  Through  the  point  b  draw  a  line  h  i  perpendic- 
ular to  a  b,  cutting  the  line  e  /in  the  point  7i,  and  the  line  d  g  in  the  point  i.  For  the 
sake  of  simplicity  we  will  now  consider  the  plane  to  be  bounded  at  the  ends  by  the 
straight  lines  e  d  and  h  i  in  place  of  the  arcs  e  a  d  and  /  b  f).  Join  the  points  d  and  h 
by  a  straight  line  d  h,  bisect  this  line — that  is,  find  the  point  j  midway  between  the 
points  d  and  7t;  join  the  points,/  and  e,  also  the  points,/  and  •?,  by  the  straight  lines  ej 
andj  i.  Divide  the  line  e  j  into  three  equal  parts,  and  thus  obtain  the  point  A-,  which 
is  the  first  point  of  division  from  the  line  d  Ji.  Also  divide  the  line  j  i  into  three  equal 
parts,  thereby  obtaining  the  point  I,  which  is  the  first  point  of  division  from  the 
line  d  Ji.  Join  the  points  k  and,/  by  a  straight  line  k  J  cutting  the  line  a  b  in  the  point 
C;  this  point  C  will  be  the  center  of  gravity  of  the  plane  e  d  h  i — that  is,  the  center  of 
gravity  of  the  plane  bounded  by  the  straight  lines  e  d,  h  i,  e  h,  and  d  i ;  but  it  will 
not  be  the  exact  center  of  gravity  of  the  plane  bounded  at  the  ends  by  the  arcs  e  a  d 
and  fig',  for  all  practical  purposes  we  may  consider  the  point  C  to  be  the  center 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


231 


of  gravity  of  the  plane  bounded  by  the  arcs,  as  in  this  case  the  error  will  not  amount 
to  more  than  i  inch — that  is  to  say,  the  distance  between  the  point  a  in  the  arc  e  a  g 
and  the  center  of  gravity  C  found  by  the  foregoing  construction,  will  only  be  £  inch 
greater  than  the  distance  between  a  and  the  true  center  of  gravity  of  the  plane 
bounded  by  the  arc  end  and/ b  g. 

To  prove  that  the  point  C  is  the  correct  center  of  gravity  of  the  plane  e  d  hi,  it 
may  be  stated  that  the  line  d  h  divides  the  plane  into  two  triangles,  d  e  h  and  d  h  i; 
the  point  k  is  the  center  of  gravity  of  the  triangle  d  e  h,  and  the  point  /  is  the  center 
of  gravity  of  the  triangle  d  h  i.  We  may  now  consider  the  two  triangles  to  form  a 
system  of  bodies ;  under  these  conditions  the  point  about  which  the  two  triangles  will 
balance  each  other  must  lie  in  a  line  joining  the  centers  of  gravity  k  and  I;  but  the 
two  triangles  make  up  the  plane  d  e  h  /,  and  we  have  seen  that  the  center  of  gravity 
of  this  plane  must  lie  in  the  center  line  a  &,  therefore  its  center  of  gravity  C  must  be 
the  point  in  which  the  lines  a  b  and  k  I  intersect. 

To  show  that  the  method  of  finding  the  centers  of  gravity  A;  and  I  of  the  triangles 
is  correct,  we  have  the  following  demonstration,  taken  from  "  Theoretical  Mechanics," 
by  J.  Weisbach. 

In  a  triangle  d  c  h,  Fig.  376,  every  line  drawn  from  an  angle  to  the  center  of  the 
opposite  side  will  contain  the  center  of  gravity  of  the  triangle.  Thus  the  line  e  m 
drawn  from  the  angle  e  to  the  center  in  of  the  opposite  side  d  h  will  contain  the  center 
of  gravity,  because  the  line  e  m  bisects  all  lines  such  as  o  p,  r  s,  which  are  drawn 
parallel  to  the  side  d  h.  The  line  d  n  drawn  from 
the-  angle  d  to  the  center  n  of  the  opposite  side  c  h 
will  also  contain  the  center  of  gravity  of  the  triangle, 
because  the  line  d  n  will  bisect  every  line  drawn  par- 
allel to  the  side  c  h,  and  therefore  the  point  of  inter- 
section k  of  the  two  lines  e  m  and  d  n  must  be  the 
center  of  gravity  of  the  whole  triangle. 

Join  the  points  n  and  m  by  a  straight  line ;  this 
line  m  n  must  be  parallel  to  the  side  e  d,  because  the 
line  m  n  is  drawn  from  the  center  of  the  side  e  h  to 
the  center  of  the  side  d  Ji.  Again,  the  length  of  the 
line  n  in  must  be  equal  to  one-half  of  the  length  of  the  side  d  e,  because  the  line  n  m  is 
drawn  from  the  center  n  of  the  side  c  h  parallel  to  e  d.  Since  n  m  is  parallel  to  d  e, 
it  follows  that  the  triangles  •;«  n  k  and  d  e  k  are  similar ;  and  because  m  n  is  equal 
to  one-half  of  e  h,  the  line  m  k  is  equal  to  one-half  of  the  line  e  k,  and  consequently 
the  line  k  m  must  be  equal  to  one-third  of  the  whole  line  e  m.  Hence  the  center  of 
gravity  k  of  the  triangle  <l  <•  h  is  at  a  distance  equal  to  J  e  m  from  the  middle  point  m 
of  the  side  d  li,  and  at  a  distance  equal  to  §  e  m  from  the  angle  e. 


Fig.  376 


ri;.\(TICAL    METHOD. 


•_>.">!).  Before  the  pradiral  method  for  determining  the  center  of  gravity  is 
explained,  let  us  first  obtain  an  insight  into  an  important  property  of  gravity.  Let 
the  outline  in  Fig.  .'!77  represent  the  shape  of  an  iron  plate  of  equal  thickness  through- 


232 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


out,  and  let  A,  B,  C,  and  I)  represent  holes  drilled  anywhere  near  the  edges  of  this 
plate.  Assume  now  that  from  a  pin  driven  into  a  wall  the  plate  is  suspended  by  the 
hole  A,  the  diameter  of  the  pin  being  a  little  less  than  that  of  the  hole,  so  that  the 
plate  can  freely  turn  on  the  pin.  We  now  find  that  after  the  plate  has  made  a  few 
oscillations,  it  will  come  to  rest  in  only  one  position ;  even  if  we  withdraw  it  from  this 
position,  the  plate,  as  soon  as  it  is  free  to  move,  wih1  again  come  to  rest  in  the  same 
place.  Let  us  carefully  mark  this  position ;  to  do  so,  we  suspend  a  chalked  line  /*  i, 
and  plummet  from  the  pin;  then,  when  the  plate  is  at  rest,  we  carefully  snap  the 
string  against  the  plate  and  thus  obtain  on  the  plate  a  chalked  mark  which  represents 
a  vertical  line  drawn  through  the  center  of  the  hole  A.  We  now  remove  the  plummet 
and  suspend  the  plate  by  the  hole  B,  then  replacing  the  plummet  we  draw  on  the 
plate  in  the  same  manner  as  before  another  vertical  chalked  line  through  the  center 
of  the  hole  B.  In  a  similar  manner  we  suspend  the  plate,  in  succession,  from  the 
holes  C  and  D,  and  from  the  center  of  these  holes  draw  vertical  chalked  lines,  thereby 
obtaining  four  chalked  lines,  which  are  represented  in  the  figure  by  broken  lines 
drawn  through  the  centers  of  the  holes  A,  B,  C,  and  D.  We  now  find  this  remarkable 


Fig.  877 


condition,  namely,  all  these  lines  will  intersect  in  one  and  the  same  point  G.  If  ad- 
ditional holes,  E,  F,  and  //,  are  drilled  anywhere  in  the  plate — near  its  edge — and 
the  plate  suspended  from  these  holes  in  succession,  and  chalked  lines  drawn  in  a 
manner  as  described,  we  will  find  that  these  additional  lines  through  the  centers  of  the 
holes  E,  F,  and  H  will  also  pass  through  the  same  point  G.  This  point  is  the  center 
of  gravity  of  the  plate.  Hence  we  may  say  that  a  vertical  line  drawn  through  the 
point  from  which  the  plate  is  suspended  will  pass  through  or  contain  the  center  of 
gravity  of  the  plate.  The  lines  which  pass  through  the  center  of  gravity  of  a  body 
are  called  "  lines  of  gravity." 

In  precisely  the  same  manner  we  can  find  the  position  of  the  center  of  gravity 


LOCOMOTIVE   COXSTRTCTIOX. 


233 


of  the  weight  shown  in  Fig.  374,  thus :  Cut  a  templet  to  the  form  of  the  surface  shown 
in  Fig.  374.  When  the  templet  is  to  be  cut  to  full  size,  it  will  be  best  to  make 
it  of  wood;  its  thickness  must  be  the  same  throughout;  indeed,  this  condition  is 
very  important;  the  real  thickness  makes  no  difference;  it  may  either  be  £  or  J  of  an 
inch,  or  1  inch,  but  it  must  be  uniform.  When  the  templet  is  to  be  cut  to  a  scale, 
it  may  be  made  of  stiff  paper ;  in  this  case  care  must  be  taken  to  keep  it  perfectly 
flat. 

On  this  templet  draw  the  center  line  a  b ;  this  line  will  contain  the  center  of 
gravity.  Anywhere  in  and  near  one  of  the  corners  of  the  templet,  punch  a  small 
smooth  hole,  and  from  a  pin  suspend  the  templet  by  the  hole,  as  shown  in  Fig.  378. 
In  doing  so  care  must  be  taken  to  allow  the  templet  perfect  freedom  to  turn  on  the 
pin.  From  the  same  pin  suspend  a  plummet  line,  and  when  the  templet  is  at  rest 
mark  off  the  point  C  in  which  the  plummet  line  intersects  the  center  line  a  b ;  this 
point  C  will  be  the  center  of  gravity  of  the  surface  shown  in  the  figure,  and  will  give 
the  distance  between  the  center  of  gravity  of  the  weight  and  the  point  b,  which  is 
equal  to  4£  inches.  Therefore  the  distance  between  the  center  of  axle  and  the  center 
of  gravity  of  this  counterbalance,  which  was  to  be  found,  is  equal  to  11  +  4£  =  15 £ 
inches.  Or,  we  may  say,  the  arm  of  the  counterbalance  is  15J  inches  long. 

260.  In  order  to  become  familiar  with  the  principles  given  in  the  previous  articles, 
we  will  apply  them  to  a  simple  case. 

EXAMPLE  78. — Let  a  in  Fig.  379  represent  the  center  of  the  axle,  b  the  center  of 
the  crank-pin;  the  distance  between  these  two  centers  is  12  inches;  ijkl  represents 


the  cast-iron  crank,  which  is  of  uniform  thickness.  The  weight  of  the  crank  is  250 
pounds.  In  addition  to  this  weight  of  the  crank,  another  weight  111  of  100  pounds  is 
applied  to  the  center  b  of  the  crank-pin.  One  view  of  the  couutei'balance  is  shown  by 
the  outline  e  d  fh,  and  all  the  dimensions  of  the  counterbalance  except  its  thickness 
are  given.  The  edge//*  of  the  counterbalance  is  to  be  placed  11  inches  from  the 
center  a  of  the  axle;  the  center  line  op,  and  the  center  line  a  b  of  the  crank,  are  to 
lie  in  one  straight  line  o  b.  It  is  required  to  find  the  weight  of  this  counterbalance, 
which  will  hold  in  equilibrium  the  sum  of  the  weight  of  the  crank  and  the  additional 
weight  of  100  pounds  applied  to  the  center  b  of  the  crank-pin.  It  is  also  required  to 


234  MODERN  LOCOMOTIVE   CONSTRUCTION. 

find  the  thickness  of  the  counterbalance.  Our  first  step  will  be  to  find  the  center  of 
gravity  of  the  crank  and  also  that  of  the  counterbalance. 

In  order  to  avoid  hereafter  a  misunderstanding,  we  deem  the  following  remarks 
necessary.  In  these  calculations,  and  in  fact  for  all  calculations  employed  for  deter- 
mining the  counterbalance  in  locomotive  wheels,  we  only  need  to  know  the  positions 
of  the  centers  of  gravity — that  is  to  say,  we  need  only  to  know  the  distances  of  the 
centers  of  gravity  of  the  crank  and  the  counterbalance  from  the  center  of  the  axle. 
Hence,  for  the  sake  of  brevity,  we  shall  speak  of  the  center  of  gravity  as  if  it  was 
located  in  the  surface  of  the  crank  or  counterbalance,  whereas  in  reality  it  lies  in  the 
center  of  the  thickness.  Thus,  in  saying  that  G  is  the  center  of  gravity  of  the 
crank,  it  must  be  understood  that  the  point  G  simply  indicates  the  distance  of  the 
center  of  gravity  from  the  center  of  the  axle,  or  the  position  it  occupies  between  the 
center  a  and  I. 

To  find  the  center  of  gravity  of  the  crank,  cut  out  a  templet  to  the  shape  of  the 
crank,  and  on  this  templet  draw  the  center  line  a  b.  Anywhere  near  the  edge  of  the 
templet  punch  a  small  smooth  hole  w,  and  then  suspend  the  templet  by  the  hole  n  from 
a  pin,  as  shown  in  Fig.  377,  allowing  it  to  have  freedom  to  oscillate.  From  the  same 
pin  suspend  a  plummet  line  and  mark  off  the  point  G  in  which  the  plummet  line 
intersects  the  center  line  a  I ;  this  point  G  will  be  the  center  of  gravity  of  the  crank. 
Suppose,  now,  that  by  this  method  we  have  found  the  distance  between  the  center  of 
gravity  G  and  the  center  a  of  the  axle  to  be  3  £  inches. 

To  find  the  center  of  gravity  of  the  counterbalance  c  dfh,  we  proceed  in  a  manner 
as  shown  in  Fig.  378,  and  explained  in  Art.  259.  Also  assume  that  by  this  method 
we  have  found  the  distance  between  the  center  of  gravity  C  and  the  edge/ h  to  be 
4£  inches. 

Since  the  distance  between  the  edge  fh  and  the  center  a  of  the  axle  has  been 
given,  namely,  11  inches,  it  follows  that  the  total  distance  between  the  center  a  and 
the  center  of  gravity  C  must  be  equal  to  11"  +  4£"  =  15£  inches. 

Our  next  step  will  be  to  find  the  weight,  which,  when  applied  to  the  center  b  of 
the  crank-pin,  will  have  the  same  effect  in  producing  rotation  of  the  crank  around  the 
center  a  as  that  of  the  weight  of  the  crank  applied  to  its  center  of  gravity  G.  This 
weight  is  found  by  Eule  43,  given  in  Art.  257.  Hence  we  have, 

250  pounds  x  3j  inches 

— =7^ — ; —  —  =  72.91  pounds. 

12  inches 

This  weight  of  72.91  pounds  applied  to  the  center  b  will  have  the  same  tendency 
to  produce  rotation  as  the  weight  of  the  crank  (250  pounds)  applied  to  the  center  of 
gravity  G,  and  therefore  the  72.91  pounds  applied  at  b  will  require  the  same  counter- 
balance as  250  pounds  applied  at  G. 

But  to  the  center  b  of  the  crank-pin  is  to  be  applied  an  additional  weight  of  100 
pounds,  represented  by  m.  Consequently,  the  total  weight  applied  to  the  crank-pin, 
and  which  must  be  counterbalanced,  is  172.91  pounds. 

We  may  now  reduce  our  problem  to  a  very  simple  one.  Thus :  Fig.  380.  Let  the 
line  c  b  represent  a  straight  lever,  and  a  its  fulcrum ;  the  distance  between  the 
fulcrum  a  and  the  end  b  of  the  lever  arm  is  equal  to  the  distance  between  the  center 


•—U)i—      — »k-- — 19—      — i\ 


MODERN  LOCOMOTIVE   CONSTRUCTION.  235 

of  axle  aiid  the  center  of  crauk-pin,  namely,  12  inches;  and  the  distance  between  the 

fulcrum  a  and  the  end  c  of  the  other  lever  arm  is  equal  to  the  distance  between  the 

(•••liter  of  the  axle  and  the  center  of  gravity 

of   the  counterbalance,  namely,   15£   inches. 

Now  we  know  that  the  weight  applied  to  the 

end  l>  is  equal  to  172.91  pounds,  consequently 

all  we  have  to  do  is  to  find  the  amount  of 

weight,  which,  when  applied  to  the  end  c,  will 

hold  the  lever  in  equilibrium.  To  find  this  weight,  or,  we  may  say,  to  find  the  weight  of 

the  counterbalance  for  any  given  weight  on  the  crank-pin,  we  have  the  following  rule : 

RVLE  44. — Multiply  the  distance  in  inches  between  the  center  of  the  axle  and  the 
center  of  the  crank-pin  by  the  total  weight  in  pounds  applied  to  the  crauk-pin,  and 
divide  this  product  by  the  distance  in  inches  between  the  center  of  the  axle  and  the 
center  of  gravity  of  the  counterbalance;  the  quotient  will  be  the  weight  in  pounds  of 
the  counterbalance. 

According  to  this  rule,  we  have, 

172.91  pounds  x  12  inches 

-  =  133.86  pounds, 
15  £  pounds 

which  is  the  weight  of  the  counterbalance. 

In  order  to  find  the  thickness  of  this  weight  or  counterbalance,  we  must  first  find 
the  number  of  cubic  inches  contained  in  it.  One  cubic  inch  of  cast-irou  is  generally 
reckoned  to  weigh  .26  of  a  pound ;  hence  the  following  rule : 

RULE  45. — Divide  the  number  of  pounds  in  the  counterbalance  by  .26;  the 
quotient  will  be  the  number  of  cubic  inches  in  the  counterbalance. 

The  number  of  pounds  in  our  counterbalance  is  133.86 ;    consequently,  according 

to  Rule  45, 

133.86 


.26 


=  514.84  cubic  inches  in  the  counterbalance. 


To  determine  the  thickness  of  the  counterbalance,  we  have  the  following  rule : 
RULE  46. — Divide  the  number  of  cubic  inches  contained  in  the  counterbalance  by 
the  area  in  square  inches  of  its  face ;  the  quotient  will  be  the  thickness  of  the  counter- 
balance in  inches. 

The  area  of  the  face  e  d  f  h  (Fig.  379)  is  found  by  multiplying  one-half  the 
sum  of  the  parallel  sides  ed  and  /  h  by  its  length  o  p;  the  product  will  be  the 
number  of  square  inches  in  the,  surface  r  </  /'//.  Thus: 

ft  -^-  3" 

x  8"  =  40  inches. 

-j 

Now,  according  to  Rule  46,  the  thickness  of  the  counterbalance  will  be, 

514.84  cubic  inches 

,      -  =  12.8  /  inches,  say  121  inches. 
40  square  inches 

261.  Generally  in  locomotives  there  is  not  sufficient  room  for  a  counterbalance  12£ 
indies  in  thickness,  consequently  we  must  employ  two  weights,  each  similar  in  form 


236 


MODERN   LOCOMOTIVE   CONSTRUCTION. 


to  that  shown  in  Fig.  379,  so  that  the  thickness  can  be  reduced  and  still  have  sufficient 
weight  for  counterbalancing. 

EXAMPLE  79. — Let  i  j  k  I,  Fig.  381,  represent  the  same  cast-iron  crank  of  250 
pounds,  and  m  the  same  additional  weight  of  100  pounds  applied  to  the  center  of  the 
crank-pin,  as  given  in  Example  78.  The  sum  of  the  weights  of  this  crank  and  weight 
m  are  now  to  be  counterbalanced  by  two  weights  whose  dimensions,  except  their 
thicknesses,  are  given,  and  these  dimensions  of  each  weight  are  the  same  as  those  of 


Fig.  381 


the  weight  given  in  Example  78.  The  space  between  the  two  weights  is  equal  to  the 
thickness  of  the  spoke  of  the  wheel ;  the  weight  of  the  spoke,  as  in  the  previous 
example,  is,  for  the  sake  of  simplicity,  left  out  of  consideration.  The  edge  fh  of  the 
upper  weight  is  to  be  placed  11  inches  from  the  center  a  of  the  axle,  and  the  same 
distance  is  to  be  maintained  between  the  edge  f.2  Ii2  of  the  lower  weight  and  the 
center  a.  In  fact,  the  only  difference  between  this  example  and  the  former  one  is  that 
two  weights  for  counterbalancing,  instead  of  one,  are  to  be  used.  It  is  required  to  find 
the  number  of  pounds  in  these  weights  and  their  thickness. 

Since  the  size  and  weight  of  the  crank,  and  also  the  additional  weight  m,  are  the 
same  as  before,  it  follows  that  the  total  weight  as  considered  to  be  applied  to  the  crank- 
pin  will  be  172.91  pounds,  and  this  weight  must  be  counterbalanced.  Again,  since  the 
size  of  each  counterbalance  weight,  except  the  thickness,  is  equal  to  the  size  of  the 
weight  shown  in  Fig.  379,  it  follows  that  the  center  of  gravity  C  of  the  upper  weight 
must  be  in  the  same  relative  position  as  before — that  is,  the  distance  between  the  center 
of  gravity  C  and  the  edge  /  h  must  be  equal  to  4i  inches.  The  same  remarks  apply  to  the 
center  of  gravity  C2  of  the  lower  weight.  Consequently  the  distance  between  the  centers 
(7 and  a  will  be  equal  to  15£  inches,  and  the  distance  between  the  center  C2  and  «  will  also 
be  equal  to  13£  inches,  both  distances  remaining  the  same  as  in  the  previous  example. 

We  may  now  consider  the  weight  e  d  fh,  and  the  weight  e.2  (I,f2  1>2  to  be  segments 
of  one  counterbalance.  Now,  considering  these  two  weights  to  form  one  counterbalance 
e  d2  f2  h,  it  must  be  apparent  that  neither  the  center  C  nor  the  center  C2  can  be  the 
center  of  gravity  of  the  whole  counterbalance  e  d.2f,  /< ;  we  must  therefore  find  a  new 
center  of  gravity  JV,  or  we  may  say  a  common  center  of  gravity  of  the  two  segments 
e  dfh  and  e.,  d2  f2  h2.  To  find  this  new  center  of  gravity,  we  simply  join  the  centers 


UODERX  LOCOMOTIVE  CONSTRUCTION.  037 

C  and  C-2  by  a  straight  line;  the  poiut  N in  which  this  line  cuts  the  center  line  r  b 
will  be  the  center  of  gravity  required.  Now  assume  that,  by  this  construction,  we  find 
the  distance  between  the  center  of  gravity  N  and  the  center  a  of  the  axle  to  be  equal 
to  15  inches;  we  have  then  all  the  data  necessary  for  determining  the  weight  and 
thickness  of  the  counterbalance. 

In  fact,  we  may  now  reduce  our  problem  to  that  of  the  simple  straight  lever,  as 
shown  in  Fig.  380,  in  which  the  line  c  b  represents  the  lever,  a  the  fulcrum,  a  b  a  lever 
arm  12  inches  long,  and  c  a  the  other  lever  arm  15  inches  long.  To  the  end  b  is 
applied  a  weight  W  of  172.91  pounds.  It  is  now  required  to  find  the  weight  li,  which, 
when  applied  to  the  end  c,  will  hold  the  lever  in  equilibrium.  It  will  be  noticed  that 
the  only  difference  between  this  problem  and  the  previous  one  represented  by  this 
figure  consists  in  the  length  of  the  lever  arm  c  a,  which  in  the  previous  problem  was 
15£  inches  long  instead  of  15  inches,  as  we  must  now  consider  it  to  be. 

Remembering  that  the  new  center  of  gravity  N  is  the  common  center  of  gravity  of 
the  two  segments,  and  is  the  only  center  of  gravity  that  can  now  enter  in  our  calcula- 
tion, we  can  find  the  sum  of  the  weights  of  the  segments  by  Eule  44.  Hence  we  have 

172.91  pounds  x  12  inches 

rrr-  -  =  138.32  pounds, 

15  pounds 

which  is  the  sum  of  the  weights  of  the  two  segments.    Consequently  the  weight  of  one 

segment  will  be 

138.32 

—^ —  =  69.16  pounds. 

a 

It  will  be  noticed  that  in  this  example  the  sum  of  the  weights  of  the  two  segments 
is  somewhat  greater  than  the  weight  of  the  counterbalance  used  in  Fig.  379,  and  yet 
in  both  examples  the  weights  applied  to  the  crank-pin,  which  were  to  be  counter- 
balanced, are  equal.  This  is  as  it  should  be,  because  in  Fig.  381  the  distance  between 
the  center  of  gravity  N  and  the  center  a  of  the  axle  is  less  than  that  between  C  and  a 
in  Fig.  379. 

We  have  found  that  the  weight  of  one  segment  of  the  counterbalance  in  Fig.  381 
is  equal  to  69.16  pounds.  In  order  to  find  the  thickness  of  one  of  these  segments, 
we  must  first  find  by  Rule  45  the  number  of  cubic  inches  contained  in  each 
segment. 

Hence  we  have, 

nr.    =  266  cubic  inches. 

Now,  since  we  know  the  area  of  the  face  c  d  fh  of  one  segment  is  equal  to  40 
square  inches,  the  thickness  of  the  segment  will  be,  according  to  Rule  46,  equal  to 

266  cubic  inches 

=  6.65  inches. 


40  square  inches  area 

262.  In  Art.  261  we  explained  the  manner  of  finding  the  common  center  of  gravity 
of  two  segments  of  a  counterbalance.  When  more  than  two  segments  are  to  be  used — 
a  frequent  occurrence — the  common  center  of  gravity  may  be  found  by  the  following 
methods : 


238 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


EXAMPLE  79a. — Let  A,  B,  D  in  Fig.  382  represent  three  segments  of  a  counter- 
balance ;  all  these  segments  are  equal  in  form  and  weight ;  it  is  required  to  find  the 
common  center  of  gravity  of  these  segments. 

RULE  47. — Draw  the  three  segments  in  their  correct  position — that  is,  leaving  the 
correct  spaces  for  the  spokes  between  them — and  draw  the  center  lines  e  f,  c2  /.„  e3  f3 ; 
these  center  lines  will,  when  produced,  pass  through  the  center  of  axle.  Find  the  center 
of  gravity  of  one  of  the  segments,  say  of  the  segment  B,  by  the  method  shown  in  Fig. 
374,  or  by  the  method  shown  in  Fig.  378,  and  thus  obtain  the  center  of  gravity  C2. 
From  the  center  a  of  the  axle  draw  an  arc  passing  through  the  center  C2,  cutting  the 
line  e/in  the  point  C,  and  the  line  e3  f3  in  the  point  C3.  Then  C  will  be  the  center 
of  gravity  of  the  segment  A ;  and  C3  the  center  of  gravity  of  the  segment  D.  The 
centers  C,  C2,  C3  must  lie  in  an  arc  described  from  the  center  of  the  axle,  because  the 
segments  are  all  placed  at  equal  distances  from  the  center  of  axle,  and  are  alike  in  form 
and  weight.  Through  the  points  C  and  C3  draw  a  straight  line  cutting  the  center  line 
e2  f2  in  the  point  i.  Divide  the  distance  between  the  center  of  gravity  C2  and  the  point 


Fig.  382 


Fig.  383 


Fig.  384 


i  into  three  equal  parts ;  the  point  of  division  G,  which  is  nearest  to  the  point  i,  will  be 
the  common  center  of  gravity  of  the  three  segments.  Here  we  see  that  the  distance 
between  G  and  i  is  equal  to  one-third  of  C2  i. 

EXAMPLE  79b. — Let  A,  J5,  D,  E  in  Fig.  383  represent  four  segments  of  a  counter- 
balance ;  all  these  segments  are  equal  in  form  and  weight ;  it  is  required  to  find  the 
common  center  of  gravity  of  these  segments. 

RULE  48. — Draw  the  segments  in  their  correct  position,  leaving  the  exact  amount 
of  space  for  the  spokes  between  them.  Two  of  these  segments  must  lie  above  the  line 
b  «,  and  two  of  the  segments  below  it.  The  line  I  a  must  pass  through  the  center  of 
the  space  for  the  wheel  spoke,  and  when  produced  pass  through  the  center  a  of  the 
axle  and  also  through  the  center  of  the  crank-pin.  Draw  the  center  lines  ef,e2f2,  e3f3, 
and  e4/4.  Find  the  center  of  gravity  of  one  of  the  weights,  say  C  of  the  segment  A, 
according  to  the  method  shown  in  Fig.  374  or  Fig.  378.  From  the  center  a  of  the  axle 
describe  an  arc  passing  through  the  center  of  gravity  C  and  cutting  the  line  e2  f2  in 


MODKKX  LOCOXOTirE   COXSTRCCTIOX.  239 

flic  point  ('.,,  c-1  ,/3  in  the  point  (73,  and  c4ft  in  the  point  64.  The  point  C2  will  bo 
the  center  of  gravity  of  the  segment  B,  C3  the  center  of  gravity  of  Z),  and  C4  that  of 
the  segment  E.  Through  the  centers  C  and  6'4  draw  a  straight  line,  cutting  the  line 
a  b  in  the  point  /;  ulso  through  the  centers  C2  and  G3  draw  a  straight  line,  cutting  «  b 
in  the  point  It ;  the  point  G  midway  between  i  and  h  will  be  the  common  center  of 
gravity  of  the  four  segments. 

EXAMPLE  79r. — Let  A,  B,  Z>,  E,  F  in  Fig.  384  represent  five  segments  of  a 
counterbalance,  all  of  them  equal  in  form  and  weight ;  it  is  required  to  find  the 
common  center  of  gravity  of  these  segments. 

EULE  49. — Through  the  center  a  of  the  axle  draw  the  horizontal  line  a  b ;  on  this 
line  draw  the  segment  D,  making  its  center  line  e3  fj  coincide  with  a  I.  Draw  the 
segments  A  and  B  above,  and  the  segments  E  and  F  below,  the  segment  D — all  in  the 
correct  position,  with  the  correct  spaces  for  the  spokes  between  them.  Find  the  center 
of  gravity  C3  of  the  segment  D,  according  to  the  method  shown  in  Fig.  374  or  Fig.  378. 
From  the  center  a  of  the  axle  describe  an  arc  passing  through  C3  and  cutting  the  center 
lines  e  f,  <:2  ./o,  P4  ,/4,  >':>  f:>  of  the  segments  A,  B,  E,  F  in  the  points  C,  C2,  C4,  C5 ;  these 
points  will  be  the  centers  of  gravity  of  their  respective  segments.  Join  the  points  Cand 
C'5  by  a  straight  line,  cutting  a  b  in  the  point  i.  Also  join  the  points  C2  C4  by  a  straight 
line,  cutting  a  I  in  the  point  h.  On  the  line  a  b  lay  off  a  point  /  midway  between  i  and 
/<;  and  then  divide  the  distance  between  I  and  G3  into  five  equal 
parts.  The  point  of  division  marked  G  nearest  to  the  point  /  will 
be  the  common  center  of  gravity  of  the  five  segments. 

The  common  center  of  gravity  G  of  the  five  segments  may  also 
be  found  approximately,  but  often  near  enough  for  practical  pur- 
poses, in  the  following  manner : 

Cut  a  templet  conforming  with  the  outline  I  m  fa  n  o  of  all  the 
five  segments  shown  in  Fig.  384. 

This  templet  is  represented  on  a  smaller  scale  in  Fig.  385.  Any- 
where in  this  templet  punch  a  small  smooth  hole  b  and  suspend  it  by 
this  hole  from  a  pin,  allowing  it  freedom  to  oscillate.  From  this  pin 
suspend  a  plummet-line;  the  point  G  in  which  the  plummet-line 
crosses  the  center  line  i  b,  previously  drawn  on  the  templet,  is  the  center  of  gravity 
of  the  five  segments. 

263.  A  knowledge  of  the  principles  upon  which  the  foregoing  methods  are  based 
may  not  only  prevent  mistakes  in  the  applications  of  the  methods,  but  will  also  enable 
us  to  find  the  common  center  of  gravity  of  a  number  of  segments  under  varied 
conditions. 

In  Fig.  386  we  may  consider  the  horizontal  line  G  to  d  to  represent  a  lever,  whose 
fulcrum  is  at  a — that  is,  the  center  of  the  axle.  This  lever  is  held  in  equilibrium  by 
the  weight  TFapplied  at  d,  and  the  counterweights  A  and  B  applied  at  the  end  of  the  ot  her 
lever  arm.  The  segments  A  and  I>  are  equal  in  form  and  weight,  and  are  placed  at 
equal  distances  from  the  center  a  of  the  axle,  consequently  their  centers  of  gravity,  C 

and  (_'.#  must  also  be  at  equal  distances  from  tl enter  «.     Again,  since  the  distance 

Itetween  the  center  of  gravity  (7  and  tin-  line  //  <\  is  equal  to  that  between  the  center 
G'2  and  the  line  b  (/,  it  follows  that  the  straight  line  C  C.,  joining  tin-  centers  of  gravity 


240  MODERN  LOCOMOTIVE   CONSTRUCTION. 

of  the  two  segments  must  be  perpendicular  to  b  d.  From  statements  made  in  previous 
articles,  we  may  assume  that  the  whole  weight  of  the  segment  A  is  concentrated  at 
the  point  C,  and  if  this  point  C  is  left  free  to  move,  the  force  of  gravity  will  cause  it 
to  move  in  the  straight  line  C  C2.  Also  the  whole  weight  of  the  segment  B  may  be 
considered  to  be  concentrated  at  its  center  of  gravity  C'2,  and  if  this  point  is  left  free 


w 


fig.  387 

Fiff.  386 

to  move,  the  force  of  gravity  will  cause  it  to  move  in  the  same  line  C  (72  prolonged. 
Therefore  we  may  say  that  the  line  C  C2  is  the  line  of  direction  in  which  the  forces 
duo  to  the  weights  of  the  two  segments  act;  and  according  to  the  remarks  in 
Art.  '235,  the  line  G  a,  which  is  the  perpendicular  distance  from  the  fulcrum  a  to  the 
line  of  direction  C  C2,  is  the  length  of  the  arm,  and  we  may  consider  the 
segments  A  and  B  to  be  directly  applied  to  the  point  G ;  or,  in  other  words,  we  may 
assume  that  the  sum  of  the  weights  of  the  two  segments  is  concentrated  at  the  point 
G,  producing  precisely  the  same  effect  in  holding  the  lever  in  equilibrium  as  the 
combined  effect  of  the  weight  of  A  acting  at  C,  and  the  weight  of  B  acting  at  C2. 

264.  But  we  may  consider  this  problem  in  another  light.  Treating  the  two 
segments  A  and  B  as  two  distinct  bodies,  similar  to  those  shown  in  Fig.  387,  each  one 
equal  to  any  given  weight — that  is  to  say,  they  may  be,  or  may  not  be,  equal  in 
weight,  placed  in  any  given  position,  either  one  above  the  other,  or  side  by  side,  or 
otherwise,  the  common  center  of  gravity  of  these  bodies  or  weights  can  be  found  in 
the  following  manner : 

Fig.  387.  Let  E  represent  one  body  weighing  10  pounds,  and  W  another  body 
weighing  5  pounds ;  it  is  required  to  find  the  common  center  of  gravity  of  these  two 
bodies. 

RULE  50. — First  find  the  center  of  gravity  c  of  the  weight  E,  and  also  the  center 
of  gravity  d  of  the  weight  W;  join  the  points  c  and  c?  by  a  straight  line ;  this  line  will 
contain  the  common  center  of  gravity  G  of  the  two  weights  E  and  W.  Now,  in  order  to 
find  the  exact  location  of  the  point  G  on  the  line  c  d,  we  have  the  following  proportion : 

The  sum  of  the  weights  E  and  W  :  the  line  c  d  :  :  weight  of  E  :  G  d; 

or, 
The  sum  of  the  weights  E  and  W  :  the  line  c  d  :  :  weight  of  W  :  G  c. 

Now,  supposing  we  find  that  by  measurement  the  line  c  d  is  12  inches  long,  then 

we  have, 

(10  +  5)  :  12  :  :  10  :  G  d. 

Working  out  this  proportion,  we  have, 

12  x  10 


M»DI:I;\  i.ocoMorii  /•:  c<>\srnrrTioy.  241 

which  shows  that  the  common  center  of  gravity  G  is  located  at  8  inches  from  d. 

Again, 

(10  +  5)  :  12  :  :  5  :  G  c. 


Working  out  this  proportion,  we  have, 

12  x  5 


15 


=  4  =  G  c, 


which  indicates  that  the  common  center  of  gravity  G  is  located  at  4  inches  from  c. 

Now  notice  the  product  of  the  weight  E  multiplied  by  its  arm  G  c  is  equal  to  the 
product  of  the  weight  W  multiplied  by  its  arm  G  d.  This  is  as  it  should  be,  otherwise 
the  solution  is  not  correct,  because  G  is  the  point  about  which  the  weights  M  and  W 
must  balance  each  other,  and  consequently  the  moment  of  the  force  R  must  be  equal 
to  the  moment  of  the  force  W.  (See  Art.  256.) 

In  Fig.  386  the  weights  of  the  two  segments  A  and  B  are  equal,  and  consequently, 
according  to  the  foregoing  rule,  the  common  center  of  gravity  G  must  lie  midway 
between  C  and  Co  in  a  straight  line  joining  these  two  points. 

The  usefulness  of  Rule  50  will  become  apparent  as  we  proceed. 

265.  Let  us  now  examine  the  method  for  finding  the  common  center  of  gravity 
of  three  segments  as  shown  in  Fig.  382. 

In  the  first  place,  let  us  assume  that  the  center  segment  B  has  been  removed ;  in 
this  way  our  problem  becomes  similar  to  that  shown  in  Fig.  386,  and  we  find  that  the 
point  i  (Fig.  382)  in  which  the  vertical  line  C  C3  cuts  the  horizontal  line  a  e2  is  the 
common  center  of  gravity  of  the  two  segments  A  and  D.  We  now  may  assume  that 
simply  a  weight  equal  to  the  sum  of  the  weights  of  the  segments  A  and  U  is  applied  at 
the  point  i,  and  throw  the  idea  of  segments  A  and  D  out  of  mind.  Replacing  the 
segment  B  in  its  proper  position,  and  remembering  that  in  determining  the  effect  of 
this  weight  we  assume  the  whole  of  the  weight  of  the  segment  is  concentrated  at  its 
r-iitcr  of  gravity  C2,  we  have  then  two  weights  applied  to  the  line  a  c2,  namely,  one  at 
( '.,  and  the  other  at  i ;  the  weight  at  i  is  twice  as  great  as  that  applied  at  C2.  Now  to 
find  the  effect  of  these  two  weights  we  must  find  their  common  center  of  gravity  G  by 
Rule  50.  For  the  sake  of  simplicity  we  will  say  the  weight  of  the  segment  B  is  equal 
to  1 ;  and  consequently  the  weight  applied  at  i  will  be  equal  to  2 ;  hence  we  have, 

The  sum  of  the  weights  at  i  and  C2  -  3  :  line  C.2  i  :  :  weight  at  C2  :  G  i. 

Working  out  this  proportion,  and  assuming  that  by  measurement  the  line  C2  i  is 
equal  to  Ij  inches,  we  have, 

H  x  1 
— o —  =  £  inch  =  G  i ; 

that  is,  the  common  center  of  gravity  G  is  located  £  inch  from  the  point  i  towards 
C2.  Now  4  inch  is  equal  to  i  of  li  inches — that  is  to  say,  G  i  =  £  of  C2  i,  which 
agrees  with  the  previous  construction. 

266.  Now  let  us  take  Fig.  383.     Here  we  have  four  segments  which  make  up  the 
counterbalance.     The  point  /'  in  which  tin-  line  C  C4  cuts  the  line  a  I  is  the  common 
center  of  gravity  of  the  two  segments  .4  and  E;  and  the  point  //  in  which  the  line 
('.,  ('3  cuts  the  line  a  b  is  the  common  center  of  gravity  of  the  two  segments  />'  and  I). 


242 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


We  may  now  assume  that  we  have  simply  a  weight  applied  at  h,  and  another  one  at  / ; 
and  since  the  weights  of  the  segments  are  equal,  the  weight  applied  at  h  is  equal  to 
that  applied  at  i.  In  order  to  determine  the  effect  of  these  two  weights  at  h  and  i,  we 
must  find  their  common  center  of  gravity  G ;  and  since  these  weights  are  equal,  we 
find,  by  Rule  50,  that  the  common  center  of  gravity  G  is  located  midway  between  /* 
and  i,  which  agrees  with  the  construction. 

267.  Lastly,  Fig.  384.  Here  we  find  the  counterbalance  composed  of  five  segments. 
The  point  i  in  which  the  line  C  C5  cuts  the  line  a  b  is  the  common  center  of  gravity  of 
the  two  segments  A  and  F.  The  point  h  in  which  the  line  C2  C\  cuts  the  line  a  b  is 
the  common  center  of  gravity  of  the  two  segments  B  and  E.  According  to  the 
construction  in  Fig.  383,  the  point  /  in  Fig.  384,  midway  between  h  and  i,  is  the 
common  center  of  gravity  of  the  weights  applied  at  li  and  i — that  is,  of  the  four 
segments  A,  -B,  E,  and  F.  We  may  now  assume  that  there  are  two  weights  applied  to 
the  line  a  I),  one  at  /  and  another  at  C3.  The  weight  at  /  is  four  times  as  heavy  as 
that  at  C3.  We  may  now  determine  the  location  of  the  common  center  of  gravity  G 
of  the  two  weights,  namely,  the  weight  at  (73,  and  that  applied  at  I  by  Eule  50 ;  hence 
we  have, 

The  sum  of  the  weights  at  I  and  C3  =  5  :  line  C3  I  :  :  weight  at  C3  :  G  I. 

Working  out  this  proportion,  and  assuming  that  by  measurement  we  find  the  line 
C3  I  to  be  equal  to  3f  inches,  we  have, 


3f  x  1 


=  £  inch  =  Gl\ 


But  f 


that  is,  the  common  center  of  gravity  is  located  at  3  inch  from  /  towards  C3. 

inch  is  i  of  3J,  which  agrees  with  the  construction. 

In  all  counterbalances  which  are  composed  of  segments,  we  always  consider  the 

sum  of  the  weights  of  the  segments  to  be  concentrated  at  their  common  center  of 

gravity,  and  the  distance  from  this  point  to  the  center  of  the  axle — that  is,  from  G  to 

a — to  be  the  lever  arm. 

268.  Figs.  388  and  389  represent  the  form  of  one  of  the  cast-iron  segments  of  a 

counterbalance  designed  to  be   bolted  between  the  spokes.     As  will  be  seen,  this 

segment  is  made  in  two  pieces,  A  and  B. 
The  piece  A  is  placed  in  the  outer  face  of 
the  driving  wheel,  and  B  in  the  inner  face 
of  the  wheel.  They  are  held  together  and 
clamped  to  the  wheel  by  the  three  bolts 
c  c  c2.  In  large  wheels  these  bolts  are  £  of 
an  inch  diameter ;  •  in  small  wheels,  $  inch 
diameter.  Since  the  space  through  which 
the  counterbalance  has  to  move  is  nearly 


Fig.  388 


Fig.  3S9 


always  limited,  the  bolt  heads  are  generally  made  conical  and  countersunk  into  the 
outer  piece  A ;  in  the  inner  piece  B  pockets  /  are  cast  to  receive  the  nuts.  The 
strips  (j  g  are  simply  chipping  strips,  so  that  the  weights  can  be  snugly  fitted  between 
the  spokes  and  rim. 


MODERX  LOCOMOTII'K   COSSTBUCTIOX.  243 

The  thickness  of  the  outer  piece  A  depends  upon  the  available  space  through 
which — in  connection  with  the  axle — it  can  revolve.  In  some  cases,  the  thickness 
of  this  piece  depends  upon  the  amount  of  weight  required.  The  inner  piece  B  is 
usually  kept  even  with  the  spokes.  The  length  d  e  also  depends  upon  the  amount  of 
weight  required,  and  varies  from  about  one-half  to  three-quarters  of  the  distance 
between  the  rim  and  hub  of  the  wheel.  These  segments  should  never  fill  the  whole 
space  from  rim  to  hub,  because,  if  these  spaces  are  closed  by  the  segments,  it  will, 
when  the  wheels  are  in  certain  position,  be  very  difficult  to  oil  the  axle  journals. 

In  calculating  the  weight  of  the  counterbalance  it  is  always  best  to  establish  first 
the  length  d  e  of  the  segment,  and  then  find  its  thickness. 

EXAMPLE  80. — It  is  required  to  find  the  dimensions  of  the  segments  of  a  coun- 
terbalance for  an  eight-wheeled  locomotive — that  is,  an  engine  having  four  driving 
wheels  and  a  four-wheeled  truck,  such  as  is  shown  in  Fig.  1.  Cylinders,  18  inches 
in  diameter;  stroke",  24  inches;  weight  of  crosshead,  154  pounds;  piston  and  rod, 
complete,  306  pounds;  main-rod,  280  pounds;  side-rod,  240  pounds;  crank-pin,  60 
pounds.  The  form  of  wheel  is  shown  in  Fig.  390. 

Not  only  the  revolving  parts,  but  also  the  reciprocating  parts  of  a  locomotive 
will  have  a  disturbing  influence  on  the  smooth  running  of  the  engine.  To  obtain 
a  steady  motion,  the  weight  of  all  the  parts  which  produce  a  disturbing  influence 
must  be  counterbalanced.  Engineers  agree  that  the  sum  of  the  weights  of  all  the 
revolving  parts  must  be  counterbalanced ;  but  as  to  the  proportion  of  the  weight  of  the 
reciprocating  parts  which  ought  to  be  counterbalanced,  they  are  not  unanimous.  A 
few  believe  that  only  two-thirds  of  the  weight  of  the  reciprocating  parts  should  bo 
counterbalanced.  Our  practice  has  been  to  counterbalance  the  sum  of  the  weights  of 
all  the  reciprocating  parts,  and  we  believe  it  is  safe  to  say  that  this  is  the  practice 
of  the  majority  of  engineers,  and  will  give  the  best  results;  we  will  follow  this 
practice  in  the  example  under  consideration. 

In  determining  the  dimensions  of  the  counterbalance,  we  need  to  confine  our 
attention  to  only  one  side  of  the  engine,  and  consequently  in  our  calculation  we  have 
to  deal  with  only  two  driving  wheels  through  which  the  counterbalance  must  be 
distributed.  Each  wheel  should  have  enough  weight  to  counterbalance  one-half  of  the 
weight  of  the  reciprocating  parts,  in  addition  to  the  total  weight  of  the  parts  which 
revolve  with  the  wheel.  Hence  our  first  step  will  be  to  separate  the  weights  of  the 
revolving  and  the  reciprocating  parts. 

The  parts,  and  their  weights,  which  revolve  with  the  main  driving  wheel,  are : 

Crank-pin 60  Ibs. 

Weight  of  crank  referred  to  pin 180    " 

One-half  side-rod 120    " 

One-half  mam-rod 140   " 

Total  weight  of  the  revolving  p.-n-ts .100  Ibs. 

The  parts,  and  their  weights,  which  revolve  with  the  rear  driving  wheel,  are: 

Crank-pin 60  Ibs. 

Weight  of  crank  jvfenvd  to  pin 180    " 

One-half  of  side-rod 120   " 

Total  weight  of  revolving  parts 360  Ibs. 


244  MODERN  LOCOMOTIVE   CONSTRUCTION. 

By  the  term  weight  of  the  crank  referred  to  the  crank-pin  is  meant  that  weight 
which,  when  applied  to  the  crank-pin,  will  have  the  same  effect  in  tending  to  turn  the 
axle  as  the  whole  of  the  weight  of  the  crank  applied  to  the  center  of  gravity  of  the 
crank,  as  explained  in  Art.  257.  To  find  the  weight  of  the  crank  referred  to  the  crank- 
pin,  we  must  first  find  the  center  of  gravity  of  the  crank  by  the  method  shown  in  Fig. 
377,  and  then  compute  the  weight  by  Eule  43.  Of  course,  in  this  case,  the  thickness 
of  the  crank  is  supposed  to  be  uniform.  When  the  thickness  is  not  uniform,  another 
method  is  often  adopted,  which  will  be  explained  in  Art.  270. 

Eeciprocating  parts,  and  their  weights,  are : 

Crosshead,  pin,  etc 154  Ibs. 

Piston  and  rod 306   " 

One-half  of  main-rod 140   '' 

Total  weight  of  reciprocating  parts 600  Ibs. 

One-half  of  this  weight  must  be  counterbalanced  in  the  main  wheel,  the  other 
half  in  the  rear  wheel.  Consequently,  the  total  weight  to  be  counterbalanced  in  the 

main  driving  wheel  will  be 

500  +  300  =  800  pounds; 

and  the  weight  to  be  counterbalanced  in  the  rear  driving  wheel  will  be 

360  -I-  300  =  660  pounds. 

We  will  first  determine  the  dimensions  of  the  segments  composing  the  counter- 
balance in  the  main  driving  wheel.  In  this  wheel  we  have  to  counterbalance  800 
pounds  applied  to  the  crank-pin,  and  since  the  distance  between  the  center  of  crank- 
pin  and  center  of  axle  is  12  inches,  the  moment  of  the  force  produced  by  the  800 

P°UndsiS  •  800X12  =  9600. 

Our  next  step  will  be  to  establish  the  number  of  segments  in  the  counterbalance, 
and  also  the  length  d  e  of  each.  No  regular  rules  for  determining  the  number  of 
segments,  and  their  lengths,  can  be  given.  An  experienced  engineer  will  readily  see 
that  four  segments,  A,  B,  Z),  and  E  (Pig.  390),  will  be  necessary.  A  smaller  number 
of  segments  would  make  their  thickness  too  great  for  the  available  space  through 
which  they  have  to  move.  Hence  we  will  decide  that  four  segments  are  to  be  used, 
and  that  the  length  d  e  of  each  one  is  to  be  12  inches. 

We  now  find  the  center  of  gravity  C  of  one  of  the  segments  in  the  manner  shown 
in  Fig.  374  or  378,  and  then  find  the  common  center  of  gravity  G  of  the  four  segments 
by  the  method  shown  in  Fig.  383  and  explained  in  Art.  266.  Measuring  the  distance 
between  the  common  center  of  gravity  G  and  the  center  a  of  the  axle,  Fig.  390,  we 
find  it  to  be  18f  inches.  Now,  since  the  moment  of  force  due  to  the  weight  of  the 
four  segments,  that  is,  the  product  obtained  by  multiplying  the  total  weight  of  the 
four  segments  by  the  18f  inches,  must  be  equal  to  the  moment  of  the  force  due  to 
the  weight  applied  to  the  crank-pin,  namely  9,600  (previously  determined),  we  can  find 
the  total  weight  of  the  four  segments  by  Rule  44.  Hence  we  have 

800  x  12        9600 
~Ia75~  =  18.75 


MODEBX  LOCOMOTIVE   COXSTRUCTIOy. 


245 


Since  the  weight  of  a  cubic  inch  of  cast-iron  is  .26  of  a  pound,  the  total  number 
of  cubic  inches  in  the  four  segments  will  be,  according  to  Eule  45, 

512 

-^  =  1969  cubic  inches. 

.*io 

Assuming  that  in  Fig.  390  the  area  of  fg  h  i  of  one  segment  is  80  square  inches, 
then  the  area  of  the  four  segments  will  be 

80  x  4  =  320  square  inches. 

And  lastly  the  thickness  of  the  counterbalance,  or,  which  amounts  to  the  same 
tiling,  the  thickness  of  each  segment,  will  be,  according  to  Rule  46, 

1969 


320 


=  6.15+  inches. 


The  dimensions  of  the  segments  for  the  rear  driving  wheel  are  found  in  a  manner 


fig.  39O 

precisely  similar  to  the  foregoing — that  is,  we  first  establish  the  length  d  e  of  the 
segments,  and  then  find  the  thickness.  We  will  make  the  calculations  without  explan- 
atory remarks. 


240  MODERN  LOCOMOTIVE   CONSTRUCTION. 

Let  us  decide  that  the  length  d  e  of  each  one  of  the  four  segments  in  the  rear 
driving  wheel  is  to  be  12  inches — that  is,  equal  to  the  length  of  segments  in  the  main 
driving  wheel. 

Under  these  conditions  the  common  center  of  gravity  G  will  be  18£  inches  from 
the  center  a  of  the  axle,  the  same  as  in  the  main  driving  wheel.  Now,  remembering 
that  the  total  weight  applied  to  the  crank-pin  in  this  wheel,  and  which  must  be 
counterbalanced,  is  660  pounds,  we  have, 

660  x  12 

1875      =  422.4  pounds, 

which  is  the  total  number  of  pounds  in  the  four  segments.    Again, 

422.4 


.26 
in  the  four  segments,  and  lastly, 


=  1624  cubic  inches 


1624 

-  u  inches, 


which  is  the  thickness  of  each  segment  in  the  rear  driving  wheel. 

269.  It  must  be  remarked  here,  that  when  the  center  of  gravity  C  of  one  of  the 
segments  is  to  be  found  by  the  method  shown  in  Fig.  378,  the  templet  must  be  cut 
somewhat  smaller  than  the  face  /  g  h  i  of  the  segment  (Fig.  390),  to  allow  for  the 
amount  of  metal  cut  out  of  the  segment  for  the  spokes  and  rim.  The  reduction  in  the 
size  of  templet  should  be  made  in  the  length  d  e,  bringing  the  arc/i  closer  to  the  arc 
g  /«;  the  reduction  should  be  made  at  the  end/i,  none  at  g  h,  because  the  end  g  h  is 
not  to  fit  any  part  of  the  wheel,  and  consequently  there  is  no  metal  cut  out  at 
this  end.  A  small  reduction  should  also  be  made  in  the  width,  bringing  the  two  sides 
fg  and  h  i  closer  together.  The  amount  of  reduction  is  generally  governed  by  good 
judgment  rather  than  by  calculation,  as  the  latter  method  is  tedious  and  involves 
considerable  labor.  Any  slight  inaccuracy  which  may  result  can  easily  be  corrected 
by  finding  the  center  of  gravity  of  the  pattern  of  the  segment  when  made,  which  will 
enable  us  to  correct  any  small  error  in  the  weight  of  the  counterbalance  by  readjusting 
the  common  center  of  gravity  G  of  all  the  segments,  and  make  such  slight  changes  in 
the  pattern  as  may  be  deemed  necessary.  In  many  cases  the  required  reduction  of  the 
templet  will  be  so  small  that  for  practical  purposes  we  may  cut  the  templet  to  conform 
to  the  surface  fg  h  i  without  making  any  reduction. 

It  must  also  be  noticed  that  the  weight  of  the  counterbalance,  found  in  the 
manner  as  we  have  done,  will  be  slightly — very  slightly — too  heavy,  for  the  following 
reason : 

If  the  wheel  center  is  made  without  the  crank-pin  hub,  and  if  the  workmanship  is 
absolutely  perfect,  the  wheel  center  itself  will  be  perfectly  balanced.  In  putting  in 
the  crank-pin  hub,  a  portion  F  of  the  spokes  must  be  cut  out,  and  for  this  amount  of 
metal  thus  cut  out  no  allowance  has  been  made,  and  consequently  the  counterbalance 
is  not  only  sufficient  for  the  weight  of  the  crank,  but  also  for  the  amount  F  cut  out  of 
the  spokes,  and  therefore  the  counterbalance  is  slightly  too  heavy,  but  by  an  amount 
barely  appreciable. 


MODERX  LOCOMOTIVE   COXSTKUCTrOX. 


247 


Fly.  391 


270.  Again,  engineers  and  draftsmen  often  find  it  difficult  to  determine  the  exact 
location  of  the  center  of  gravity  of  the  crank,  and  therefore,  instead  of  finding  the 
weight  of  the  crank  referred  to   the  crank-pin,  content  themselves   by  finding  the 
weight  of  the  crank-pin  hub — that  is,  the  weight  of  the  metal  around  the  crank-pin 
as  indicated  by  the  shaded  portion  in  Fig.  391 — and  then  adding  this  weight  to  the 
revolving  paiis  to  be  counterbalanced,  in 

place  of  the  weight  of  the  crank  referred 
to  the  crank-pin ;  by  this  method  the  neces- 
sity  of  finding  the  center  of  gravity  of  the 
crank  is  avoided. 

This  method  of  adding  the  weight  of 
the  crank-pin  hub  to  the  revolving  parts,     _"        ~ 
which  must  be  counterbalanced,  saves  con-  J 

siderable  labor  in  all  cases,  and  is  usually 
adopted  when  the  crank  is  not  of  uniform 
thickness. 

271.  Frequently,  in  fact  in  the  majority 
of  locomotives,  the  counterbalance  and  the 
wheel  center  are  cast  in  one  piece,  as  shown 

in  Fig.  392.  To  determine  the  dimensions  of  a  counterbalance  of  this  kind,  we  should 
follow  the  rales  given  in  Art.  268.  That  is  to  say,  we  should  consider  the  solid  coun- 
terbalance to  be  made  up  of  a  number  of  segments  bolted  between  the  spokes,  as  shown 
in  Fig.  390,  and  then  determine  the  thickness  by  a  method  precisely  similar  to  that 
given  in  Art.  268. 

The  correctness  of  the  result  of  the  calculation  depends  upon  the  correct  position 
of  the  common  center  of  gravity  G  (Fig.  390)  of  the  segments,  and  this  center,  of 
course,  depends  upon  the  correct  position  of  the  center  of  gravity  C  of  each  segment ; 
therefore  we  conclude  that  when  the  centers  of  gravity  C\  C2,  etc.,  of  the  segments  are 
placed  in  incorrect  positions,  the  results  of  our  calculations  will  be  erroneous. 

If  in  these  segments  metal  had  not  to  be  cut  out  of  the  sides  i  h  and///  to  fit  the 
spokes,  and  also  out  of  the  end  /  i  to  fit  the  rim,  we  could  find  very  accurately  and 
easily  the  centers  of  gravity  Cl  C2  of  these  segments  of  unifonn  thickness  by  the  method 
shown  in  Fig.  378.  But  it  is  not  such  an  easy  matter  to  find  the  center  of  gravity  of 
segments  of  irregular  shapes,  as  shown  in  Fig.  390,  or  Fig.  389;  in  fact,  to  find 
correctly  the  center  of  gravity  of  one  of  these  segments  involves  a  great  amount  of 
labor,  unless  we  have  the  pattern  of  which  the  center  of  gravity  may  be  found  by 
balancing  it  on  a  knife  edge. 

Now,  when  a  solid  counterbalance,  as  shown  in  Fig.  392,  is  considered  to  be  made 
up  of  segments,  we  cannot  lose  sight  of  the  fact  that  the  sides  of  the  segments  and 
their  outer  ends  must  be  formed  to  fit  the  spokes,  and  consequently  we  have  to  do 
with  segments  of  irregular  shape.  To  reduce  the  labor  of  finding  the  common  center 
of  gravity  of  these  segments,  or,  in  other  words,  to  reduce  the  labor  of  finding  the 
center  of  gravity  G  of  the  counterbalance  shown  in  Fig.  392,  we  may  generally  adopt 
the  following  method;  the  results  obtained  by  this  method,  although  not  absolutely 
correct,  will  be  sufficiently  accurate  for  practical  purposes: 


248 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


Instead  of  considering  the  counterbalance  (Fig.  392)  to  be  made  up  of  segments, 
as  we  should  do,  we  treat  it  as  it  appears  to  be,  namely,  as  one  solid  weight.  Now, 
assume  that  the  dimensions  of  this  weight  are  such  as  will  give  a  sufficient  amount  of 
metal,  and  not  more,  to  counterbalance  all  the  weights  applied  to  the  center  of  crank- 
pin  ;  then,  by  inserting,  so  to  speak,  this  counterbalance  into  the  wheel  center,  we  must 
cut  portions  out  of  several  spokes,  and  also  cut  a  part  out  of  the  rim ;  by  so  doing  we 
leave  unbalanced  portions  of  other  spokes  and  a  part  of  the  rim  on  the  crank-pin 
side,  all  of  which  are  precisely  similar  in  amount  and  form  to  those  cut  out.  It  must 
be  evident  that  by  leaving  portions  of  the  wheel  itself  unbalanced  we  create  new- 
disturbing  forces,  which  will  be  just  as  injurious  as  a  similar  amount  of  unbalanced 
weight  applied  to  the  crank-pin.  Therefore  the  thickness  of  our  present  counter- 
balance must  be  increased,  so  that  not  only  the  weights  applied  to  the  crank -pin  are 
counterbalanced,  but  also  have  sufficient  weight  to  counterbalance  those  portions  of 
the  wheel  opposite  those  which  had  to  be  cut  out  to  make  room  for  the  counter- 
balance. In  the  following  example  it  will  be  seen  how  a  close  approximation  to  the 
exact  thickness  can  be  determined. 

EXAMPLE  81. — The  distance  from  the  center  of  axle  to  the  center  of  crank-pin  is 
12  inches ;  the  total  weight  applied  to  the  crank,  which  must  be  counterbalanced,  is 


L 

Fig.  393 

Section  through  r.s 


Fig.  392 


500  pounds ;  the  depth  of  the  counterbalance  from  h  to  g  is  to  be  8  inches,  and  is  to 
terminate  in  the  centers  a  b,fe  of  the  spokes  A  and  B  (Fig.  392).  It  is  required  to 
find  the  thickness  of  the  counterbalance ;  this  thickness  is  indicated  by  t,  in  Fig.  393. 
This  figure  represents  a  cross-section  of  the  counterbalance  through  the  line  marked 
r  s,  in  Fig.  392. 

In  the  first  place,  make  a  drawing  of  the  wheel,  as  shown  in  Fig.  392,  then  cut  a 
templet  conforming  to  the  outline  a  I  hfeg  of  the  counterbalance,  and  find  its  center 
of  gravity  G  in  the  manner  shown  in  Fig.  385.  Mark  off  this  center  of  gravity  G  on 
the  drawing,  and  measure  its  distance  from  the  center  i  of  the  axle.  Assume  that  we 
find  the  distance  between  G  and  the  center  i  to  be  16  inches,  we  will  then  have  all 
the  data  necessary  for  detei'mining  the  thickness  by  calculation.  The  moment  of  the 
force  due  to  the  weight  applied  to  the  crank-pin  is  equal  to 

500  x  12  =  6000. 


MODERN  LOCOMOTIVE  CONSTRUCTION.  249 

The  moment  of  the  force  due  to  the  weight  of  the  counterbalance  must  also  be 
equal  to  6,000  ;  hence,  by  dividing  6,000  by  the  distance  of  the  center  of  gravity  G 

from  the  center  /,  we  have, 

6000 

—Tfr  =  375  pounds, 

which  is  the  weight  necessary  to  counterbalance  the  weight  applied  to  the  crank-pin. 
But  this  counterbalance  cuts  portions  out  of  five  spokes  and  also  a  portion  out  of  the 
rim.  Suppose  now  that  by  calculation  we  find  the  total  amount  thus  cut  out  of  the 
wheel  center  to  be  equal  to  80  pounds.  Adding  this  weight  to  375  pounds  previously 

found,  we  have, 

375  +  80  =  455  pounds 

for  the  total  weight  of  the  counterbalance. 

We  now  find  the  thickness  of  the  counterbalance  in  a  manner  precisely  similar  to 
that  given  in  Art.  268,  thus  :  Dividing  the  total  weight  of  455  pounds  by  the  weight 
of  a  cubic  inch  of  cast-iron,  we  have, 

455 

~n~  =  1750  cubic  inches 

required  in  the  counterbalance. 

And  lastly,  dividing  the  number  of  cubic  inches  by  the  area  of  the  face  a  I  It/eg 

of  the  counterbalance,  we  have, 

1750 

•~-  =  6.83  inches, 


which  is  the  thickness  of  the  counterbalance. 

272.  The  simplest  way  of  finding  the  area  of  the  surface  a  b  life  g  of  the  counter- 
balance will  be  to  describe  an  arc  m  n  o  midway  between  the  arcs  age  and  b  hf(not 
drawn  through  the  center  of  gravity  <?),  then  find  by  measm-ement  the  length  of  the 
arc  >n  n  o,  and  multiply  this  length  by  the  depth  a  b  ;  the  product  will  be  the  number 
of  square  inches  in  the  face  of  the  counterbalance. 

273.  Now  it  must  not  be  understood  that  we  claim  to  obtain,  by  the  foregoing 
1  1  n  -t  hod,  an  absolutely  correct  thickness  for  the  counterbalance;  the  thickness  thus 
found,  although  close  enough  for  practical  purposes,  is  only  approximate,  for  the 
following  reason  :  In  the  first  place,  we  should  have  considered  the  counterbalance  to 
consist  of  a  number  of  segments  fitted  around  the  spokes  and  part  of  the  rim,  and 
then  found  the  common  center  of  gravity  of  these  segments.     Instead  of  this,  we  found 
the  center  of  gravity  of  the  whole  counterbalance  without  making  the  proper  allow- 
ances for  the  amount  of  metal  which  must  be  cut  out  of  the  counterbalance  for  the 
spokes  and  rim. 

Consequently  the  center  of  gravity  G,  as  we  have  determined  it,  will  be  somewhat 
—  a  very  small  amount  —  too  far  away  from  the  center  i  of  the  axle.  Secondly,  simply 
adding  the  weight  of  those  portions  of  tlit-  sjiokt-s  and  rim  which  are  cut  out  for  the 
purpose  of  making  room  for  the  counterbalance  to  the  weight  required  for  coun- 
1i-rli;ilaiifing  the  weights  applied  to  the  crank-pin,  is  not  quite  correct.  The  exact 
procedure  would  be  to  find  the  common  center  of  gravity  of  the  portions  cut  out  of  tin- 


250 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


wheel  center,  so  as  to  obtain  the  correct  distance  between  this  center  of  gravity  and 
the  center  of  axle,  aiid  then  find  the  amount  of  weight  which,  when  applied  to  the 
center  of  gravity  G  of  the  counterbalance,  will  have  the  same  effect  as  the  weight  of 
the  portions  of  the  spokes  and  rim,  cut  out  of  the  wheel  center,  applied  to  their  own 
center  of  gravity,  and  then  add  this  weight,  so  found,  to  the  weight  required  for 
counterbalancing  the  weights  applied  to  the  crank-pin.  But  such  a  course  requires  a 
great  amount  of  labor,  and  may,  on  account  of  its  complicacy,  and  the  errors  which 
may  creep  into  it,  give  no  better  results  than  those  obtained  by  the  simpler  method. 

274.  In  ten-wheeled,  Mogul,  and  consolidation  engines,  the  weight  of  the  recipro- 
cating parts  which  are  to  be  counterbalanced  is  often  equally  distributed  throughout 
the  driving  wheels,  thereby  making  the  counterbalances  in  all  wheels,  excepting  the 
main  wheel,  equal  in  size.     The  counterbalance  in  the  main  wheel  will,  of  course,  be  a 
little  larger  than  the  others,  because  it  has  to  counterbalance  a  heavier  crank-pin  than 
those  in  the  other  wheels,  and  has  also  to  counterbalance  the  weight  of  one-half  of  the 
connecting-rod,  which  the  counterbalances  in  the  other  wheels  have  not  to  do.     In 
some  instances  this  arrangement  will  require  in  the  main  wheel  a  counterbalance 
whose  thickness  is  too  great  to  pass  through  the  available  space;  consequently,  in 
such  cases,  the  total  weights  of  the  revolving  and  reciprocating  parts  are  assumed  to 
be  equally  distributed  throughout  all  the  wheels,  and  calculations  of  the  counter- 
balance made  accordingly,  making  the  counterbalances  in  all  the  wheels  (main  wheel 
included)  equal  in  size  and  weight. 

275.  Even  with  this  arrangement,  considerable  difficulty  is  often  experienced  in 
the  endeavor  to    obtain  sufficient  weight   in  the  counterbalance  for  narrow-gauge 
locomotives  having   driving  wheels  of    comparatively   small  diameter,   say  3  feet. 
In  fact,  in  many  of  these  small  wheels  it  is  impossible  to  obtain  sufficient  weight 
with  cast-iron,  and  therefore  these  counterbalances  are  frequently  cast  hollow  and 
filled  with  lead — for  instance,  such  as  is  shown  in  Fig.  394.     This  figure  represents 


Section  throityh  a  b 
Fif/.  396 


Fiff.  394 


a  wheel  center  33  inches  diameter,  and  is  designed  for  a  narrow-gauge  (3  feet,  or 
3  feet  6  inches  gauge)  locomotive.  Fig.  395  represents  a  section  of  the  counter- 
balance through  the  line  a  b  (Fig.  394),  and,  as  will  be  seen,  it  is  cast  hollow,  so  that 
it  may  be  filled  with  lead.  We  use  lead  because  it  is  considerably  heavier  than  cast- 
iron,  and  consequently  we  can  obtain  a  counterbalance  which  will  need  less  room 


MODERN  LOCOMOTIVE   CONSTRUCTION.  251 

thaii  a  oast-iron  counterbalance  of  the  same  weight.  We  are  compelled  to  use  a 
counterbalance  of  small  dimensions,  and  yet  a  heavy  one,  because,  on  account  of  the 
small  diameter  of  the  wheel,  the  length  of  the  arc  x  x2  %  (which  is  the  length  of  the 
counterbalance)  and  also  the  depth  a  y  are  limited.  The  available  space  through 
which  the  counterbalance  has  to  move  will  also  limit  the  thickness  d2  e2  in  Fig.  395. 
Another  fact  we  must  not  lose  sight  of,  is  that,  occasionally,  in  small  wheels,  the 
distance  between  the  center  of  gravity  G  of  the  counterbalance  and  the  center  of  the 
axle  will  be  less  than  the  length  of  the  crank ;  consequently  in  cases  of  this  kind  the 
weight  of  the  counterbalance  will  be  greater  than  that  applied  to  the  crank-pin ;  such 
conditions  do  not  occur  in  large  wheels.  Small  wheels  leave  us  but  very  little  choice 
in  the  length  x  x2  x3  of  the  counterbalance,  which  frequently  must  be  extended  around 
the  wheel  as  far  as  we  can  possibly  go  with  advantage.  The  choice  of  the  width  a  y 
is  also  limited ;  all  we  can  do  is  to  make  it  as  wide  as  we  can,  leaving  only  sufficient 
room  between  the  counterbalance  and  the  hub,  that  is,  between  a  and  p,  for  oiling  the 
axle  journals  when  the  counterbalance  stands  above  the  center  j. 

276.  The  thickness  of  this  class  of  counterbalances  can  be  determined  by  the 
following  method,  which  saves  labor  and  promotes  simplicity,  but  will  give  only 
approximate  results,  though  "close  enough  for  practical  purposes. 

EXAMPLE  82. — The  length  of  the  crank  is  9  inches;  total  weight  applied  to  the 
crank-pin,  which  is  to  be  counterbalanced,  is  300  pounds ;  it  is  required  to  find  the 
thickness  of  the  counterbalance. 

Make  a  drawing  of  the  wheel  center  as  shown  in  Fig.  394.  Lay  in  the  counter- 
balance, and  let  it  extend  from  the  center  I  m  of  the  arm  A  to  the  center  o  n  of  the 
arm  B.  This  length  of  the  counterbalance  is  arbitrary ;  an  experienced  designer  will 
know  that,  under  the  conditions,  it  must  be  made  as  long  as  possible ;  and  anything 
beyond  the  lines  I  m  and  o  n  will  add  little,  if  any,  appreciable  effect  to  the  counter- 
balance. Let  us  also  decide  to  make  the  width  a  y  equal  to  6  inches;  this  will  leave 
about  as  little  room  between  a  and  p  as  we  can  get  along  with  in  oiling  the  axle 
journal.  Let  us  now  consider  that  our  counterbalance  is  simply  a  cast-iron  box 
whose  cross-section  is  rectaugulai-,  as  represented  by  the  lines  d2f,,  f2  p2,  p2  e^  and  d2  e2, 
in  Fig.  395 ;  also  let  us  consider  that  the  cross-section  of  the  lead  is  represented  by  the 
rectangle  d  efg  1i.  Now  the  depth  d2f2  of  the  box  is  established,  the  thickness  of  the 
sides  of  this  box  is  also  established,  which  is  to  be  £  of  an  inch.  We  may,  therefore, 
find  at  once  the  weight  of  sides  d2  f2  and  e2  p.2  in  the  following  manner :  Midway 
between  the  arcs  m  a  o  and  I  y  n  draw,  from  the  center,/,  the  arc  x  X2  x3;  multiply  the 
length  of  this  arc  by  the  depth  a  y,  by  the  thickness  of  the  side,  and  by  the  weight  of  a 
cubic  inch  of  cast-iron ;  the  product  will  be  the  weight  required ;  thus : 

Assume  that  by  measurement  we  find  the  length  of  the  arc  x  X2  X3  to  be  29 
inches;  then  29"  x  G"  x  3"  x  .26  =  33.93  pounds,  which  is  the  weight  of  one  side, 
and  33.93  x  2  =  67.86  pounds,  which  is  the  weight  of  both  sides,  d.2f.2  and  e2p2. 

Let  us  now  assume  that  tho  counterbalance  is  divided  into  a  number  of  slabs,  as 
indicated  by  the  dotted  lines  </.,/„  ,/4/4,  etc.,  each  slab  1  inch  thick,  tin-  divisions  or 
cutting  planes  being  parallel  to  the  face  I  m  a  y  o  n  of  the  counterbalance.  Let 
us  now  find  the  weight  of  one  of  these  slabs.  Each  one  of  them  is  composed  of 
two  kinds  of  metal,  namely,  lead  and  cast-iron.  Of  course,  lead  predominates.  The 


252  MODERN  LOCOMOTIVE   COXSTBVCTIOJf. 

weight  of  lead  is  generally  reckoned  at  0.41  pound  per  cubic  inch.  Consequently  the 
area  bounded  by  the  dotted  lines  r  s,  t  u,  and  the  arcs  s  v  t,  r  iv  u,  multiplied  by  .41, 
will  give  the  number  of  pounds  of  lead  in  one  slab.  The  area  of  this  surface  of  lead 
is  found  by  multiplying  the  length  of  the  arc  #4  £2  %  by  width  v  w.  The  width  v  ic  is, 
under  the  given  conditions,  equal  to  4£  inches ;  and  assuming  that  by  measurement 
we  find  the  length  of  the  arc  #4  x2  X5  equal  to  27J  inches,  we  have  for  the  weight  of 
lead  in  one  slice,  whose  thickness  is  1  inch,  27£"  x  4£"  x  .41  =  50.7375  pounds,  say 
50IJ  pounds. 

To  this  weight  must  be  added  the  weight  of  the  cast-iron  included  by  the  arcs 
m  a  o  and  s  v  t ;  also  that  between  the  arcs  r  w  u  and  I  y  n,  all  1  inch  in  depth.  Let 
us  assume  that  we  have  found  by  calculation  the  weight  of  this  amount  of  cast-iron  to 
be  equal  to  10£  pounds,  then  the  total  weight  of  one  slab  will  be  equal  to 

503  +  10i  =  61  pounds. 

Cut  a  templet  to  conform  to  the  outline  I  m  a  o  n  y  of  the  counterbalance,  and  find 
its  center  of  gravity  G  by  the  method  shown  in  Fig.  385 ;  lay  off  this  center  of  gravity 
on  the  drawing,  and  measure  its  distance  from  the  center  j  of  the  axle ;  this  distance, 
we  will  say,  is  equal  to  8£  inches. 

The  moment  of  the  force  due  to  the  weight  applied  to  the  crank-pin  is  equal  to 
the  product  of  the  length  of  the  crank  into  the  weight  applied  to  the  pin ;  hence  we 

have 

300  x  9  =  2700. 

The  moment  of  the  force  due  to  the  weight  of  the  counterbalance  is  also  equal  to 
2,700 ;  hence  the  total  weight  of  the  counterbalance  must  be  equal  to 

2700 

-£-T  =  317.64+  pounds. 

o.O 

But  we  have  already  found  that  the  weight  of  the  two  cast-iron  sides  d2  f2  and 
e2  p.2  (Fig.  395)  is  equal  to  67.86  pounds.  Subtracting  this  weight  from  the  total  weight 
of  the  counterbalance,  we  have 

317.64  —  67.86  =  249.78  pounds. 

This  weight  of  249.78  pounds  must  now  be  made  up  by  the  weight  of  the  slabs 
into  which  the  counterbalance  has  been  divided.  Therefoi'e,  we  must  now  find  the 
number  of  slabs  required  to  make  up  the  weight  of  249.78  pounds.  Since  we  have 
found  that  the  weight  of  each  slab  is  equal  to  61  pounds,  we  have 

249.78 

— ^i —  =  4.09,  say  4  slabs. 

And  lastly,  since  each  slab  is  1  inch  thick,  the  thickness  of  the  lead  from  d  to  e 
must  be  4  inches,  and  the  total  thickness  from  <72  to  e2  of  the  counterbalance  must 
be  equal  to  4  +  l£  =  5J  inches.  In  this  case  we  have  left  out  the  weight  of  the  arms 
and  rim,  which  must  necessarily  be  cut  out  of  the  wheel  to  insert  the  counterbalance, 
but  this  weight  is  generally  made  up  by  filling  the  opening  y  h  i  k  with  lead. 

277.  Figs.  396,  397  represent  a  42-inch  wheel  center  with  lead  counterbalance, 
designed  in  one  of  our  prominent  locomotive  works.  It  is  suitable  for  heavy  freight 
engines  with  cylinders  20  inches  diameter  and  up  to  22  inches. 


Mi>ltKR\   LOCOMOTIVE    COXKTRVCTION. 


253 


Figs.  398,  399  represent  a  driving  wheel  with  lead  counterbalance.  It  is  used  on 
elevated  railroads.  We  believe  that  in  this  wheel  the  arms  and  rim  could  have  been 
made  lighter,  and  still  be  strong  enough  to  do  excellent  service. 


counterbalance 

'\ 


254 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


278.  In  locomotive  construction  we  have  sometimes  to  find  the  areas   of  plane 
surfaces  or  figures  similar  to  that  shown  in  Fig.  400. 

The  area  of  such  a  figure  may  be  found  in  the  following  manner : 
Divide  the  line  a  I  into  any  number  of  equal  parts,  say  five,  and  through  the 
points  of  division  c,  fj,  /,  k,  draw  lines  perpendicular  to  a  I  terminating  in  the  curve 
c  h  d.    The  lines  ef,  g  h,  etc.,  are  called  ordinates,  and  for  the  sake  of  simplicity  we 


r 

Q 

pt. 

*•< 



4 

r-  i 

d 

I- 


, 

Fiy.  400 


Fig.  403 


may  also  consider  the  lines  a  c  and  b  d  to  be  ordinates,  although  in  reality  they  are 
bounding  lines  of  the  plane.  Let  P,  P.,,  P3,  P4,  P5,  P6  represent  the  lengths  of  the 
ordinates,  and  s  the  distance  between  any  two  successive  ordinates — that  is  to  say, 
s  =  e  g  or  g  i.  The  area  of  this  figure  may  then  be  found  by  the  following : 

RULE  51. — To  find  the  area  between  a  given  curve  and  a  straight  line,  such  as 
c  h  d  and  a  b  in  Fig.  400.  To  one-half  the  lengths  of  the  extreme  outer  ordinates  add 
the  length  of  all  the  intermediate  ordinates,  and  multiply  the  sum  by  the  distance 
between  any  two  successive  ordinates ;  the  product  will  be  the  area.  Or,  putting  this 
rule  in  the  shape  of  a  formula,  we  have 

(i  P  +  Pz  +  P^  +  P^  Pr>  +  J  P6)  x  s  =  area. 
EXAMPLE  83. — Suppose  we  find  the  length  of  a  c  equal  to  2&  inches,  ef=3,gh  = 


LOCOMOTITE  COXSTRVCTJOX.  255 

2 );:,  /  j  =  2,^.,  i-  i  =  2-jij,.,  b  d  =  2,  and  the  distance  ,s  between  any  two  successive 
ordinates  equal  to  1J  inches ;  it  is  required  to  find  the  area. 

97 //  2"\ 

-f-  4-  3"  +  2ii"  +  2-jV'  +  2iV  +  ^-j  x  1J"  =  16|  square  inches. 

The  accuracy  of  the  result  depends  on  the  number  of  ordinates ;  the  greater  the 
number,  the  greater  the  accuracy. 

In  the  next  article  we  will  give  practical  applications  of  this  rule. 

279.  In  a  number  of  locomotives  the  outer  foimm  of  the  rim  of  the  driving  wheel  is 
uniform  throughout.  Such  a  wheel  is  shown  in  Fig.  358.  There  we  notice  that  a 
portion  of  the  rim,  opposite  to  the  crank,  is  cast  solid,  and  the  remaining  portion  of 
the  rirn  is  cast  hollow.  But,  in  many  cases,  the  extra  weight  of  metal  gained  by 
casting  part  of  the  rim  solid  is  not  sufficient  for  counterbalancing  the  whole  weight 
applied  to  the  crank-pin ;  consequently  we  often  find  the  whole  rim  cast  hollow,  and 
of  equal  dimensions  throughout,  with  a  certain  portion  of  the  rim  opposite  the  crank 
filled  with  lead,  so  as  to  obtain  a  greater  weight  in  the  same  amount  of  space  than  can 
be  obtained  by  casting  a  portion  of  the  rim  solid.  Hence,  it  frequently  happens  that 
the  weight  of  the  lead  in  the  rim,  and  its  effect  as  a  counterbalance,  must  be  calculated. 

Let  Fig.  401  represent  a  part  of  a  driving  wheel  whose  rim  is  cast  hollow,  and  of 
uniform  cross-section  throughout.  The  section  around  the  center  line  b  c  represents  a 
cross-section  of  the  rim ;  and  Fig.  402  represents  the  same  section  on  a  larger  scale. 
The  portion  of  the  rim  from  the  rib  a  to  the  rib  «6  is  filled  with  lead.  It  is  required  to 
find  the  weight  of  the  lead,  and  its  effect  as  a  counterbalance;  that  is  to  say,  to 
determine  the  amount  of  weight  applied  to  the  crank-pin  which  the  lead  in  the  rim 
can  counterbalance. 

This  problem  includes  two  distinct  problems,  namely,  to  find  the  weight  of  the 
lead ;  second,  to  find  the  effect  of  the  weight  of  the  lead  as  a  counterbalance. 

For  the  sake  of  simplicity  we  shall  leave  the  ribs  «3,  «4,  a6  which  are  cast  in  the  rim 
out  of  consideration,  and  proceed  as  if  the  hollow  part  of  the  rim  from  the  rib  a  to  the 
rib  a6  had  been  a  clear  space  and  then  filled  with  lead. 

WEIGHT  OF  THE  LEAD  IN  THE  BIM. 

In  looking  at  this  counterbalance,  or  body  of  lead,  from  a  theoretical  standpoint, 
we  may  consider  this  body  or  solid  to  be  generated  by  a  surface  which  is  equal  in 
extent  to  the  cross-section  of  lead  revolving  about  the  center  of  the  axle,  and  this 
surface  we  may  call  the  generating  surface.  The  contents  (that  is,  the  number  of 
cubic  inches)  of  this  solid  is  obtained  by  multiplying  the  area  of  the  generating  surface 
in  square  inches  by  the  length  of  the  path  (or  arc,  in  this  case)  in  inches  described  by 
the  center  of  gravity  of  the  generating  surface ;  the  product  will  be  the  number  of 
cubic  inches  in  the  cotmterbalance.  And  lastly,  the  number  of  cubic  inches  multi- 
plied by  the  weight  of  one  cubic  inch  of  metal  will  be  the  weight  of  the  counterbalance. 
Consequently  we  have  the  following : 

RULE  52. — Multiply  the  area  in  square  inches  of  the  cross-section  of  the  lead  by 
the  length  of  the  arc  in  inches,  described  by  the  center  of  gravity  of  this  cross-section  ; 
the  product  will  be  the  number  of  cubic  inches  in  the  counterbalance.  This  last 


256  MODERN  LOCOMOTIVE   CONSTRUCTION. 

product,  multiplied  by  the  weight  of  one  cubic  inch  of  lead,  namely,  .41  pound,  will  be 
the  weight  of  the  counterbalance. 

From  the  foregoing  we  see  that  in  determining  the  number  of  cubic  inches  in  the 
counterbalance  two  distinct  factors  enter  into  our  calculation,  namely,  the  area  of  the 
cross-section  of  the  lead,  and  the  length  of  path  which  the  center  of  gravity  of  this 
section  will  describe. 

The  outline  of  the  cross-section  of  the  lead  is  represented  by  the  outline 
d  e  h  c  n  kj  of  the  cavity  in  Fig.  402.  The  area  of  this  cross-section  can  be  found  by 
Rule  51.  Thus :  Draw  the  line  I  c  perpendicular  to  p  r.  If  the  outline  of  the  section 
is  symmetrical,  the  line  b  c  must  be  drawn  through  the  center  of  the  section ;  if  the 
outline  of  this  section  is  not  symmetrical,  we  first  draw  a  line  s  t  (any  length)  parallel 
to  line  p  r,  and  touching  the  curve  i  c  o  in  one  point,  as  c,  and  then  through  this  point, 
that  is,  the  point  of  tangency,  draw  the  line  b  c  perpendicular  to  p  r,  as  before. 

Let  us  assume  that  the  outline  of  the  section  is  symmetrical.  Through  the  point 
e  draw  a  line  e  ez  perpendicular  to  the  line  I  c.  Divide  the  distance  from  e.,  to  c  into 
any  number  of  equal  parts,  say  five ;  through  the  points  of  division  f2,  g2,  h.2j  and  i2 
draw  lines  perpendicular  to  b  c,  cutting  the  outline  of  the  section  in  the  points  /,  c/,  h,  i. 
Suppose  now  that  we  find,  by  measurement,  the  line  e  e2  to  be  equal  to  f  inch ;  ff2  = 
1& !  9  9i  =  1 5  '*  lh  =  i  5  *  *2  =  8 ;  and  c  of  course  =  0 ;  also,  the  distance,  such  as 
f2  g.2,  or  g2  h2,  that  is,  the  distance  between  any  two  successive  ordinates  or  perpendicu- 
lars, equal  to  ^6-  inch.  The  area  of  the  surface  e2  c  life  will  then,  according  to  Rule 
51,  be  equal  to 

3"  \ 

~  +  14"  +  1"  +  I"  +'  S"  +  f  ]  X  J\,  =  2.32+  square  inches. 

To  this  we  must  add  the  area  of  the  rectangle  b  d  e  e.2 ;  and  since  b  e2  is  equal  to 
£  inch,  and  e  e2  =  £  inch,  we  have  .875  x  75  =  .65+  square  inch;  hence  2.32  +  .65  = 
2.97  square  inches,  which  is  the  area  of  the  section  above  the  line  b  c ;  and  2.97  x  2  = 
5.9  square  inches,  the  total  area  of  the  cross-section  of  the  lead. 

If  the  outline  of  the  section  is  not  symmetrical,  as  we  have  assumed  it  to  be,  then 
we  find  area  of  the  surface  above  the  line  b  c,  and  that  below  the  line  b  c,  each  one 
separately,  in  a  manner  precisely  similar  to  the  foregoing,  and  add  the  two  together ; 
the  sum  will  be  the  total  area  of  the  surface. 

Our  next  step  will  be  to  find  the  length  of  the  path  described  by  the  center  of 
gravity  of  the  cross-section  of  the  lead.  Cut  a  templet  to  the  outline  d  c  h  c  n  kj  of 
the  cavity  (Fig.  402),  and  find  its  center  of  gravity  G  by  the  method  shown  in  Fig. 
378,  and  mark  this  point  G  in  its  correct  position  from  b  in  the  section  shown  in  Fig. 
401.  From  the  center  of  the  axle  describe  an  arc  passing  through  the  center  G,  and 
terminating  in  the  sides  of  the  brackets  a  and  a&.  This  arc  will  repi'esent  the  path  of 
the  center  of  gravity  G  of  the  cross-section,  while  this  cross-section,  or  generating 
surface,  is  revolving  about  the  center  of  the  axle.  Now  suppose  that,  by  measurement, 
we  find  the  length  of  this  arc  from  a  to  a6  to  be  57  inches,  then  the  total  number  of 
cubic  inches  in  the  lead  will  be  equal  to  the  product  obtained  by  multiplying  the  area 
in  square  inches  of  the  cross-section,  previously  found,  by  the  length  in  inches  of  the 
arc  a  «4  a6 ;  hence  we  have  5.9"  x  57"  =  336.3  cubic  inches.  Multiplying  the  cubic 


VODER*  LOCOMOTIl'K   COXSTRCCTJOX.  257 

inches  by  the  weight  of  one  cubic  inch  of  lead,  we  have  336.3  x  .41  =  137.883  pounds, 
which  is  the  total  weight  of  the  lead  in  the  i'im. 

EFFECT  OF  THE  LEAD  COUNTERBALANCE  IN  THE  RIM. 

280.  Our  next  step  will  be  to  determine  how  much  of  the  weight  applied  to  the 
crank-pin  will  be  counterbalanced  by  the  weight  of  the  lead  in  the  rim.  To  do  this  we 
must  find  the  center  of  gravity  G.,  (Fig.  401)  of  the  whole  amount  of  lead,  and  the 
correct  distance  between  this  center  of  gravity  and  the  center  of  the  axle.  But  here  a 
little  difficulty  arises  ;  we  see  that  the  thickness  of  the  lead  is  not  uniform,  that  is  to 
say,  the  lines  k  e,  If,  m  g,  etc.,  in  the  cross-section  of  the  lead,  as  shown  in  Fig.  402,  are 
not  equal  in  length  ;  therefore  we  cannot  find  the  center  of  gravity  of  the  lead  by  the 
method  shown  in  Fig.  385,  or  by  any  method  previously  given  ;  hence  we  adopt  the 
following  method  : 

In  the  first  place  we  shall  assume  that  the  area  of  the  cross-section  of  the  lead  is 
infinitely  reduced,  so  that  the  whole  counterbalance  will  be  represented  by  the  arc 
a  m  ffl6  (Fig.  401),  which  is  described  from  the  center  of  the  axle  and  passes  through 
the  center  of  gravity  G  of  the  cross-section  ;  and  we  shall  also  assume  that  the  whole 
weight  of  the  lead  is  concentrated  along  this  arc,  or,  in  other  words,  the  weight  of  this 
arc  o  m  «6  is  equal  to  the  whole  weight  of  the  lead.  Join  the  extremities  a  and  «fi  of 
this  arc  by  a  straight  line,  which  will  be  the  chord.  We  have  now  reduced  the  condi- 
tions of  the  problem  to  those  of  a  very  simple  one,  for  we  have  now  only  to  find  the 
center  of  gravity  of  the  arc  a  m  a6. 

In  J.  Weisbach's  "  Theoretical  Mechanics  "  we  find  that  the  center  of  gravity  of  an 
arc  of  a  circle  must  lie  in  the  radius  drawn  through  the  middle  of  the  arc,  because  this 
radius  is  an  axis  of  symmetry  of  the  arc  ;  consequently  in  our  problem  the  center  of 
gravity  G.2  of  the  arc  a  m  a6  must  lie  in  the  line  n  m,  which  is  drawn  through  the 
middle  m  of  the  arc  to  the  center  of  the  axle.  Again,  we  find  that  the  distance  from 
the  center  of  gravity  of  the  arc  to  the  center  from  which  the  arc  has  been  described  is 
to  the  radius  of  the  arc  as  the  chord  is  to  the  arc.  To  indicate  these  proportions  by 
symbols,  let 

Length  of  the  arc  be  represented  by  ........................................  b 

Length  of  the  chord  "  "  ........................................  s 

Length  of  radius        "           "  "  ........................................  r 

Distance  of  center  of  gravity  "  "  ........................................  y 

Then  we  have 

y  :  r  :  :  s  :  b.     Consequently, 

s  r 


From  the  foregoing  we  can  establish  the  following  rule  : 

RULE  53.  —  To  find  the  center  of  gravity  of  the  lead  in  the  rim,  multiply  the  length 
in  inches  of  the  chord  joining  the  extremities  of  an  arc  which  is  described  from  the 
ci-nter  of  the  axle  and  passing  through  the  center  of  gravity  of  the  cross-section  of  the 
lead  by  its  radius  in  inches,  and  divide  this  product  by  the  length  of  the  arc  in  inches; 


258 


MODERN  LOCOMOTIVE   CONSTRrCTIOX. 


the  quotient  will  be  the  distance  from  the  center  of  the  axle  to  the  center  of  gravity  of 
the  arc. 

We  have  already  seen  that  the  length  of  the  arc  a  m  a6  is  57  inches ;  now  suppose 
we  find  by  measurement  that  the  chord  a  aG  is  48J  inches,  and  the  radius  29£  inches, 
then  the  distance  of  the  center  of  gravity  G2  of  the  lead  in  Fig.  401  from  the  center 
of  axle  will  be  equal  to 

48.5"  x  29.5'^ 


57" 


=  25.1  inches. 


We  have  found  that  the  weight  of  the  lead  is  137.883  pounds,  consequently  the 
moment  of  the  force  due  to  the  weight  will  be  equal  to 

137.883  x  25.1  =  3460.86+. 

If  we  now  assume  that  the  length  of  the  crank  is  12  inches,  then  our  lead  will 
counterbalance  a  weight  applied  to  the  crank-pin  equal  to 


3460.86 


=  288.4  pounds. 


WEIGHT   OF   CRANK-PIN    HUB. 


281.  In  Art.  254  it  was  shown  that,  in  order  to  determine  the  amount  of  weight 
required  to  counterbalance  the  weight  of  the  crank,  we  must  first  find  the  center  of 


Fig.  4O4 

gravity  of  the  crank.  In  Art.  270  it  will  also  be  seen  that,  on  account  of  the  difficulty 
of  determining  accurately  the  center  of  gravity  of  the  crank,  many  draftsmen  and 
engineers  will  often,  instead  of  counterbalancing  the  weight  of  the  crank  referred  to 
the  crank-pin,  content  themselves  by  simply  adding  the  weight  of  the  crank-pin  hub 
to  the  other  weights  applied  to  the  crank-pin,  and  then  counterbalancing  the  sum  of 


MOHKRX  LOI-it.MOTJI'K   <;O\XTIircTI(>\.  259 

weights.     Hence  the  question  may  arise:  How  can  we  find  the  weight  of  the 
mink-pin  hub? 

The  weight  of  the  craiik-j>iu  hub  can  be  determined  by  the  same  method  and  rules 
which  were  employed  for  finding  the  weight  of  lead  in  the  rim  of  a  wheel,  as  explained 
in  Art.  279.  Thus,  for  instance :  Let  Fig.  403  represent  the  longitudinal  section  of 
the  crank-pin  hub  whose  weight  we  are  to  determine.  We  first  find  the  area  of  the 
section  of  the  metal  /  m  n  o  p.  To  do  this  we  follow  Rule  51.  Therefore,  divide  the 
distance  n  p  into  any  number  of  equal  parts,  and  through  the  points  of  division  draw 
lines  perpendicular  to  the  line  n  p,  terminating  in  the  curve  m  I  o.  Let  A,  B,  C,  etc., 
represent  the  length  in  inches  of  the  perpendicular  lines  (the  lengths  being  obtained 
by  measurement).  Then,  according  to  Rule  51, 

\-z  +  B+C  +  D  +  E  +  aj  x  s  =  area  in  square  inches. 

We  must  now  find  the  center  of  gravity  G  of  the  surface  I  m  n  o  p.  To  do  this, 
cut  a  templet  to  the  outline  of  this  surface,  and  find  its  center  of  gravity  by  the 
method  shown  in  Fig.  377.  Mark  off  this  center  of  gravity  G  on  the  section  of  the 
hub,  and  measure  its  distance  G  h  from  the  center  line  i  k ;  twice  the  distance  G  h  will 
be  equal  to  the  diameter  of  the  circle  whose  circumference  will  be  the  path  of  the 
point  G  as  the  surface  I  m  n  o  p  revolves  around  the  center  line  i  k.  According  to 
Rule  52,  the  number  of  cubic  inches  in  the  hub  will  be  equal  to  the  product  obtained 
in  multiplying  the  circumference  in  inches  of  the  circle  described  by  the  center  of 
gravity  G,  by  the  area  in  square  inches  of  the  surface  I  m  n  o  p.  This  last  product, 
multiplied  by  the  weight  of  one  cubic  inch  of  cast-iron  (.26),  will  be  the  weight  of  the 
crank-pin  hub. 

DRIVING   WHEEL  TIRES. 

282.  At  present,  nearly  all  driving  wheel  tires  are  made  of  steel.      The  most 
common  practice  is  to  shrink  the  tires  on  the  wheel  centers,  as  explained  in  Art.  248. 
When  tires  are  to  be  shrunk  on,  they  are  bored  out  to  a  uniform  diameter  throughout, 
so  that  their  sections  will  appear  like  that  shown  in  Fig.  404.     Of  course  the  inside 
diameter  of  the  tire  must  be  a  certain  amount  less  than  the  diameter  of  the  wheel 
center ;  care  must  be  taken  not  to  allow  too  much  for  shrinkage,  as  this  will  draw  the 
tire  too  tight,  and  will  be  liable  to  burst  it  when  the  engine  is  running;  on  the  other 
hand,  with  an  insufficient  allowance  for  shrinkage,  the  tire  will,  in  a  short  time, 
become  loose;  the   shrinkage  diameters  adopted  by  the  American   Railway  Master- 
Mechanics'  Association  are  given  in  Art.  248. 

283.  When  steel  tires  are  properly  shrunk  on  the  wheel  centers  in  this  manner, 
they  will  remain  tight  and  resist  all  the  lateral  thrust  caused  by  the  flanges  bearing 
against  the  rails,  until  they  are  worn  considerably.     Then  these  tires  need  watching, 
as  the  constant  rolling  action  is  liable  to  loosen  them.     When  this  is  the  case,  shim- 
ming pieces  are  often  placed  ltd  ween  the  tires  and  wheel  centers.     Occasionally  the 
excellent  plan  as  shown  at  A  in  Fig.  404  is  adopted.     Here  it  will  be  seen  that  a  slight 
projection  or  flange  is  left  on  the  periphery  of  the  wheel  center  near  its  inner  face,  or 
flanged   side  of  the  wheel.     Tin-   purpose  of  this  design   is  that  the  projection  shall 

the  thrust  caused  l.y  the  flanges  bearing  against  the  rails,  and  prevent  the 


260 


MODERN  LOCOMOTJfE   COXSTRVCTION. 


tire  from  slipping  inward  should  it  at  any  time  become  loose.  Yet  it  seems  that  the 
general  inclination  is  to  avoid  the  extra  expense  incurred  in  putting  the  tires  on  the 
wheel  centers  in  this  manner,  and  consequently  the  most  common  practice  is  to  bore 
out  the  tires  straight. 

284.  Very  seldom  do  we  find  tires  which  have  not  been  shrunk  on,  but  placed 
cold  on  the  wheel  centers ;  in  fact,  we  know  of  only  one  road  in  the  United  States 


Kim  of  Wheel  Centre 


Arm  I 


Fig, 


406 


where  this  practice  prevails.  On  this  road  the  tires  are  bored  out,  tapered,  and  the 
periphery  of  the  wheel  center  turned  to  suit,  as  shown  in  Fig.  405.  The  large 
diameter  of  the  bore  of  the  tire  is  on  the  flanged  side,  or,  in  other  words,  on  the 
inside  of  the  wheel.  The  tire  is  prevented  from  slipping  outward  by  the  hook-headed 
bolts  B.  The  taper  prevents  the  tire  from  slipping  inwards,  and  resists  the  thrust 
caused  by  the  flange  bearing  against  the  rails.  Eight  hook-headed  bolts  B  are  gen- 
erally used  for  a  tire  63  inches  inside  diameter ;  and  six  bolts  B  for  a  tire  35J  inches 
inside  diameter.  The  nuts  on  these  bolts  are  not  allowed  to  project  beyond  the  inside 
of  tire ;  consequently,  recesses  for  the  nuts  are  cast  in  the  rim  of  the  wheel  center ; 
another  view  of  one  of  the  recesses,  and  metal  partially  surrounding  it,  cast  to  the 
rim,  is  shown  in  Fig.  406. 

The  advantage  claimed  for  this  mode  of  putting  tires  on  the  wheel  centers  is  that 
they  can  be  removed  when  necessary  with  less  labor  and  delay  than  will  be  required 
for  removing  tires  which  are  shrunk  on  the  wheel  centers,  and  also,  should  the 
constant  rolling  action  loosen  the  tires  at  any  time,  they  can  be  quickly  tightened. 
We  decidedly  prefer  the  tires  shrunk  on  wheel's  center,  as  we  believe  that,  with  these, 
safety  is  promoted  and  better  results  obtained. 

285.  In  fitting  up  a  pair  of  driving  wheels,  the  wheel  centers  are  first  pressed  on 
the  axles  with  a  pressure  equal  to  that  given  in  Art.  252.  The  keys  are  then  driven, 
and  the  tires  shrunk  on.  After  this  the  wheels  and  axle  are  placed  in  a  quartering 
machine  and  the  holes  bored  for  the  crank-pins.  It  is  very  important  to  have  the 
crank-pins  exactly  in  line  with  the  axle,  and  at  equal  angles  in  all  the  different  driving 
wheels,  and  such  results  are  obtained  in  a  correctly  designed  quartering  machine. 

The  crank-pins  are  then  pressed  into  the  wheel  centers  with  a  pressure  of  about 


MODERN  LOCOMOTIVE  CONSTRUCTION.  261 

six  tons  for  every  inch  in  diameter  of  the  crank-pin  fit,  providing  the  holes  are 
perfectly  true  and  smooth.  If  the  holes  are  not  perfectly  true,  which  may  not  only  be 
due  t<>  liad  workmanship,  but  may  be,  and  sometimes  is,  the  result  of  the  bad  practice 
of  shrinking  the  tires  on  the  wheel  centers  after  the  crank-pin  holes  have  been  bored, 
a  pressure  of  nine  tons  per  inch  in  diameter  is  necessary  for  pressing  the  crank-pins 
into  the  wheels.  The  last  operation  is  to  turn  the  tires  to  the  correct  gauge. 

CLEARANCE  BETWEEN  FLANGES  AND  RAILS. 

286.  The  tires  are  placed  on  the  wheel  centers  so  as  to  allow  a  certain  amount 
of  clearance,   C  (Fig.  404),  between  the  flanges  and  the  rails.     The  most  common 
amount  of  clearance  allowed  at  C  for  each  wheel  is  f  of  an  inch,  making  the  total 
clearance  between  the  rails  and  the  flanges  of  a  pair  of  driving  wheels  £  inch.    On  a 
few  roads  this  clearance  for  each  wheel  is  only  \  inch,  and,  on  the  other  hand,  we 
occasionally  find  it  increased  to  J  inch.    We  are  in  favor  of  §  inch  clearance  for  each 
wheel.     Should  the  distance  between  the  rails  of  a  curve  be  established,  and  any  doubt 
exist  as  to  the  sufficiency  of  this  amount  of  clearance  in  running  over  the  curve,  the 
proper  amount  of  clearance  can  readily  be  obtained  by  making  a  plan  of  the  cui-ve, 
and  also  a  plan  of  the  wheels  with  the  correct  distances  between  the  axles  drawn 
upon  the  former ;  the  distance  between  the  flanges  of  a  pair  of  driving  wheels  can 
then  be  regulated  to  allow  a  necessary  amount  of  clearance,  which  will  prevent  the 
wheels  from  binding  in  running  over  the  curve. 

PLAIN  TIRES. 

287.  Fig.  407  represents  a  cross-section  of  a  tire  without  a  flange ;  tires  of  this  kind 
are  generally  called  "  plain  tires."     These  tires  are 

generally  wider  than  the  flanged  tires  under  the 
same  engine. 

The  width  of  the  plain  tires  must  be  made  to 
suit  the  curves  over  which  the  engine  has  to  run, 
and  should  be  sufficiently  wide  not  to  leave  the 
track  at  any  time.    To  determine  this  width,  a  plan    cj-r 
of  the  curve  and  wheel  base  must  be  made,  and         t—  — «"-+-  — •" 

Fig.  4O7 

the  width  of  the  plain  tire  regulated  so  as  to  cover 

the  rails  at  all  times.    As  a  consequence,  the  width  of  the  plain  tires  varies,  under  the 

different  classes  of  engines,  from  5£  to  7  inches. 

DISTRIBUTION  OF  FLANGED  TIRES   IN  THE  VARIOUS  CLASSES    OF  LOCOMOTIVES. 

288.  In  a  large  majority  of  eight-wheeled  passenger  engines,  in  which  four  of  the 
whole  number  of  wheels  are  driving  wheels,  as  shown  in  Fig.  1,  all  the  driving  wheels 
have  flanged  tires.     We  have  noticed  only  a  very  few  of  this  class  of  engines  in  which 
one  pair  of  drivers — the  front  ones — had  plain  tires,  leaving  only  one  pair  of  driving 
wheels — the  rear  ones — to  guide  the  engine  over  a  curve.     This  we  believe  to  be  dan- 
gerous practice.     All  wheels  in  this  class  of  engines  should  have  flanges. 


262  MODERN  LOCOMOTIVE   CONSTRUCTIVE. 

In  a  great  number  of  Mogul  engines,  in  which  six  wheels  of  the  whole  number  are 
driving  wheels,  as  shown  in  Fig.  2,  the  front  and  rear  driving  wheels  have  flanged 
tires ;  the  middle  pair  have  plain  tires.  We  believe  this  to  be  the  best  practice,  for 
the  following  reasons :  All  locomotives  have  a  tendency  to  sway  in  front  from  side  to 
side;  this  is  objectionable.  Now,  a  Mogul  engine  has  only  a  two-wheeled  truck 
(generally  called  a  pony  truck)  in  front,  and  such  a  truck  without  the  aid  of  the 
flanges  on  the  first  pair  of  drivers  is  not  suitable  to  guide  the  engine  steadily  on  a 
straight  track,  neither  is  it  suitable  to  guide  it  safely  over  a  curve.  In  short,  more 
than  two  wheels  with  flanges  (particularly  when  these  wheels  are  truck  wheels)  are 
needed  to  guide  the  engine  steadily  in  front. 

In  ten-wheeled  engines,  in  which  six  wheels  of  the  whole  number  are  driving 
wheels,  as  shown  in  Fig.  3,  the  rear  pair  of  drivers,  and  the  pair  next  to  it,  have 
flanged  tires.  In  these  engines  the  four-wheeled  truck  in  front  is  generally  considered 
to  be  sufficient  to  guide  the  engine .  steadily  on  a  straight  track  and  safely  over  a 
curve  without  the  aid  of  flanges  on  the  front  driving  wheels. 

In  consolidation  engines,  in  which  eight  wheels  of  the  whole  number  are  driving 
wheels,  as  shown  in  Fig.  4,  the  distribution  of  flanged  tires  somewhat  differs  on  the 
various  railroads.  All  these  engines  have  in  front  a  two-wheeled — that  is,  a  pony — 
truck,  and  since,  as  we  have  stated  before,  a  two- wheeled  truck  is  generally  considered 
insufficient  to  guide  an  engine  steadily  on  a  straight  track  or  safely  over  a  curve,  we 
find  that  nearly  every  engine  of  this  class  has  flanged  tires  on  the  front  drivers.  As 
to  the  other  driving  wheels  under  the  same  engines,  we  often  find  the  rear  pair,  and 
the  pair  next  to  it,  with  flanged  tires,  leaving  only  one  pair  of  driving  wheels  (the 
pair  next  to  the  front  ones)  with  plain  tires. 

This  arrangement  seems  to  indicate  that  some  designers  consider  two  pairs  of 
the  whole  number  of  wheels,  including  truck  wheels,  to  be  necessary  for  the  pur- 
pose of  guiding  the  rear  end  of  the  locomotive,  and  two  pairs  of  wheels  to  guide  the 
front  end. 

But  this  arrangement  is  by  no  means  a  universal  one,  as  we  frequently  meet 
with  consolidation  engines  in  which  only  the  rear  and  front  pairs  of  driving  wheels 
have  flanged  tires,  leaving  the  two  intermediate  pairs  of  driving  wheels  with  plain 
tires.  This  arrangement  seems  to  indicate  that  if  one  pair  of  wheels  with  flanges  is 
sufficient  to  guide,  as  is  generally  the  case,  the  rear  end  of  a  Mogul  engine,  one  pair  of 
wheels  with  flanges  will  also  be  sufficient  to  guide  the  rear  end  of  a  consolidation 
engine,  and  that  plain  tires  on  the  two  intermediate  driving  wheels  will  allow  the 
engine  to  curve  easier.  In  a  few  instances  we  find  the  flanged  tires  distributed  so  as 
to  bring  the  flanges  on  alternate  driving  wheels.  In  such  cases  the  rear  and  third  pair 
of  driving  wheels,  counting  from  the  rear  end,  will  have  flanged  tires,  the  second  pair 
from  the  rear  and  the  front  pair  will  have  plain  tires.  Our  favorite  arrangement  is 
to  have  flanges  on  the  front  and  rear  driving  wheels ;  these,  we  believe,  will  guide  the 
engine  steadily,  and  allow  it  to  curve  easily. 

Lastly,  occasionally  we  meet  with  consolidation,  Mogul,  and  ten-wheeled  engines 
under  which  all  the  driving  wheels  have  flanged  tires. 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


263 


DEPTH  OF  FLANGES. 

289.  The  depth  D  E  of  the  flanges  of  tires  (Fig.  408)  varies,  according  to  the 
ju< lament  of  the  designer,  from  1  to  1J  inches;  and  in  a  few  instances  the  flanges  are 
made  l£  inches  deep,  as  shown  in  Fig.  405.  Many  master-mechanics  claim  that  flanges 
1  inch  deep  are  perfectly  safe,  and  to  prove  their  assertion  point  to  engines  with 


fcsJH 


flanges  1  inch  deep  on  the  drivers  which  are  and  have  been  running  daily  without  any 
trouble  or  accidents ;  yet  others  claim  that  flanges  l£  inches  deep  are  safer,  and  adopt 
this  depth  for  all  their  driving  wheel  flanges.  In  a  large  majority  of  locomotives  the 
driving  wheel  flanges  are  l£  inches  deep — an  arithmetical  mean  of  1  and  1|. 


FORMS   OF  TREAD   ON   DRIVING  WHEELS. 

290.  At  present  various  forms  of  tread  on  driving  wheels  are  used.    Fig.  408 
represents  a  section  of  a  tire  in  which  the  tread  a  b  is  a  surface  of  uniform  taper. 
This  taper  varies  on  the  different  roads  from  £  to  |  of  an  inch  in  12  inches.     This  form 
of  tread  is  usually  called  the  cone  tread. 

Fig.  409  represents  a  section  of  a  tire  in  which  the  tread  is  composed  of  two 
surfaces;  one  surface  is  represented  by  the  line  c  rf,  and  is  turned  to  a  uniform 
diameter ;  the  other  surface  is  represented  by  line  d  e,  and  is  turned  to  a  uniform 
taper ;  in  some  cases  this  taper  is  as  much  as  J  of  an  inch  in  1J  inches.  This  form  of 
tread  is  often  called  the  straight  tread. 

Fig.  410  represents  a  section  of  tire  in  which  the  tread  is  also  composed  of  two 
surfaces,  fg  and  g  h.  The  surface  fg  is  turned  to  a  slight  taper,  as  shown ;  and  the 
surface  g  h  is  turned  to  a  greater  taper,  whose  dimensions  are  also  given.  This  form 
of  tread  is  a  modification  of  the  cone  tread,  and  its  advantages  will  be  presently 
explained. 

BEST   FORM   OF   TREAD. 

291.  The  question  now  arises :  Which  is  the  best  form  of  tread  for  locomotive 
tires  I    We  have  already  seen  that  all  locomotives  have  a  tendency  to  sway  from  side 
to  side  at  the  front  end,  and  such  a  sinuous  motion  is  objectionable.     Tires  with  a 
cone  tread,  as  shown  in  Fig.  408,  will  certainly  reduce  to  some  extent  this  objection- 
able  sinuous  motion.     We  may  therefore  claim  for  the  cone  tread,  particularly  in 
engines  which  have  only  a  pony  truck  to  guide  the  front  end,  and  where  some  depend- 
ence must  be  placed  on  the  front  drivers  to  guide,  in  some  measure,  the  engine 
steadily,  that  the  cone  tread  will  help  to  keep  the  engine  in  the  center  of  the  track, 


264  MODERN  LOCOMOTIVE   CONSTRUCTION. 

and  to  some  extent  reduce  the  wear  on  the  flanges.  True,  the  taper  of  the  tread  near 
the  base  of  the  flange  will  soon  wear  off,  and  a  channel  or  groove  in  the  tire  will  be 
formed,  as  shown  in  Fig.  411.  In  this  case  the  form  of  the  ridge  near  the  flange  will, 
to  some  extent,  answer  the  same  purpose  as  that  for  which  the  cone  tread  was 
originally  designed.  Hence  we  may  say  that  the  cone  tread  is  particularly  desirable 
for  engines  in  which  the  tires  are  new,  or  have  not  been  worn.  Indeed,  this  fact  has 
often  been  proven  by  experience  on  roads,  and  more  distinctly  so  on  roads  with  sharp 
curves,  where  straight  treads  had  to  be  abandoned  and  replaced  by  cone  treads,  so  as 
to  obtain  more  satisfactory  results,  cause  the  engine  to  run  with  greater  safety  over 
the  curves,  and  reduce  the  wear  of  the  flanges  when  new.  On  the  other  hand,  the 
claim  that  a  cone  tread  will  cause  the  engine  to  run  over  a  curve  with  less  friction 
and  less  wear  on  the  tires  than  with  treads  of  other  forms  is  a  fallacy.  For  instance, 
the  argument  that  the  centrifugal  force  will  cause  the  engine  to  run  towards  the  outer 
rails  of  a  curve,  and  thus  cause  the  outer  driving  wheels  to  work  on  their  large 
diameter  and  the  inner  driving  wheels  to  work  on  their  small  diameter,  thereby 
creating  a  rolling  motion  similar  to  that  of  a  cone  on  a  flat  surface,  cannot  be 
sustained  when  sound  principles  are  applied. 

It  is  true  that  when  a  single  cone  is  rolled  over  a  flat  surface  it  will  run  in  a 
curve  whose  radius  is  dependent  on  the  taper  of  the  cone.  But  if  we  now  take  two 
equal  cones  and  fix  their  axes  rigidly  and  parallel  to  each  other  in  a  frame,  so  as  to 
obtain  conditions  similar  to  those  of  parallel  axles  and  wheels  under  an  engine,  and 
then  roll  the  cones  over  a  flat  surface,  the  results  thus  obtained  will  greatly  differ  from 
the  results  noticed  in  rolling  one  of  the  cones  separately.  The  two  cones  thus  held 
together  will  roll  in  a  curve  which  approaches  a  straight  line ;  considerable  sliding 
friction  will  also  be  created,  so  that  the  forces  required  for  rolling  both  cones  held 
together  will  be  considerably  greater  than  the  sum  of  the  two  forces  required  for 
rolling  each  cone  separately. 

The  claim  that  the  centrifugal  force  will  cause  the  engine  to  run  towards  the 
outer  rails  of  the  curved  track  is  also  very  doubtful,  because,  to  obtain  such  a  result, 
the  radius  of  the  curve  would  have  to  be  suitable  for  the  speed  of  the  engine,  or  the 
engine  would  have  to  be  run  at  a  speed  suitable  for  the  radius  of  the  curve,  and  these 
are  conditions  which  cannot  be  obtained  in  practice.  Hence  we  conclude  that  the 
cone  tread  will  not  lessen  the  wear  or  friction  on  the  tires  in  running  over  a  curve, 
but  it  has  this  advantage,  that  it  will  generally  prevent,  to  some  extent,  the  sinuous 
motion  of  the  locomotive,  and  that  on  this  account  it  will  somewhat  lessen  the  wear 
of  the  flanges. 

The  tread  shown  in  Fig.  410,  and  treads  similar  to  it,  have  in  late  years  greatly 
grown  in  favor ;  and  we  believe  these  to  be  the  best  to  adopt,  as  they  possess  not  only 
the  advantage  of  the  cone  tread,  but,  furthermore,  possess  the  advantage  of  wearing  to 
a  better  form  than  the  cone  tread,  because  the  increased  taper  of  the  tread  near  the 
edge  of  the  tire  will  reduce  the  outer  ridge  of  the  groove  or  channel  shown  in  Fig. 
411,  the  groove  being  always  an  objectionable  feature. 

The  straight  tread  shown  in  Fig.  409  possesses  only  the  advantage  of  wearing  to 
a  form  approaching  that  of  Fig.  410,  but  it  does  not  possess  the  advantage  of  steadily 
guiding  the  engine  when  the  tires  are  new. 


MODERX  LOCOMOTirE   COXSTKVCTION. 


265 


92.  In  the  Proceedings  Twentieth  Annual  Convention,  1887,  of  the  American 
Railway  Master-Mechanics'  Association,  we  see  it  reported  that  the  form  of  the  tread, 
such  as  shown  in  Fig.  412,  has  been  adopted  as  a  standard,  which  is  the  same  form  of 

Ml 


•v 


-X-* 

^^ 

—  — 

'-;'•• 

-i    ;;- 

x4 

=  -     \  ~      ^  —  • 

-^ 

i. 

^          ** 

ir 

r^j,"1       Fig.  412 


tread  as  that  adopted  by  the  Master  Car  Builders'  Association,  and  has  been  illustrated 
in  the  report  of  their  twenty-first  annual  convention,  from  which  our  illustration  has 
been  taken.  Figs.  409,  410,  and  411  have  been  taken  from  the  report  of  the  proceed- 
ings of  the  seventeenth  annual  convention  of  the  American  Railway  Master-Mechanics 
Association. 

THE  LIMIT  OP  WEAR  OF  TIBES. 

293.  A  fixed  limit  to  which  the  thickness  of  a  tire  can  be  reduced  by  wear  without 
impairing  the  safety  of  running  the  engine  cannot  be  established,  as  the  weight  on 
the  drivers,  the  climate,  and  quality  of  the  steel  will  have  some  influence  on  the  limit 
of  wear.  But  generally,  experience  seems  to  indicate  that  in  warm  climates  tires 
made  of  good  homogeneous  steel  may  remain  in  use  until  their  thickness  has  been 
reduced  to  1J  inches ;  for  light  engines,  whose  weight  does  not  exceed  10,000  pounds  on 
each  driver,  the  tires  may  remain  in  use  until  their  thickness  has  been  reduced  to  li 
inches.  In  colder  climates,  although  sometimes  we  there  see  engines  running  with 
tires  reduced  to  1J  inches  in  thickness,  we  believe  that  safety  will  be  promoted  by 
removing  the  tires  before  this  limit  has  been  reached. 


THICKNESS  AND  WIDTH  OF  TIRE. 

294.  On  many  roads  the  thickness  of  the  tire  when  new  is  3  inches,  measured 
at  the  base  of  the  flange  as  indicated  by  the  dimension  line  in  Fig.  408.  On  some 
roads  this  thickness  has  been  considerably  increased,  so  that  we  now  find  many  loco- 
motives with  tires  4  inches  thick.  The  advantage  claimed  for  thick  tires  is  that  they 
will,  to  some  extent,  reduce  the  expense  of  keeping  the  engine  in  working  order, 
because,  generally,  tires,  of  whatever  original  thickness,  are  not  allowed  to  wear  below 
a  thickness  of  1J  inches,  and  are  then  taken  off  the  wheel  center;  consequently  the 
interval  between  the  renewals  of  tires  4  inches  thick  will  be  greater  than  the  interval 
between  the  renewals  of  3-inch  tires,  and  thereby  time  and  labor  are  saved  when  the 
former  thickness  has  been  adopted.  Attain,  the  amount  of  steel  tin-own  away,  so  to 
speak,  when  the  tire  is  condemned  on  account  of  its  reduction,  will,  in  comparison  witli 
the  original  weight  of  the  tire,  be  less  for  tires  4  inches  thick  than  for  those  whose 
original  thickness  was  below  4  inches.  On  the  other  hand,  with  the  use  of  heavy 


266  MODERN  LOCOMOTIVE   CONSTRUCTION. 

tires  a  greater  dead  weight  than  desirable,  that  is  to  say,  a  weight  not  supported  by 
the  springs,  is  placed  on  the  rails,  and  such  a  procedure  must  in  time  injure  the  rails 
It  is  therefore  probable  that  in  the  future  the  thickness  of  tires  will  not  exceed  4 
inches. 

The  width  of  flanged  tires  generally  varies  from  5£  to  of  inches. 

DISTANCE  BETWEEN  THE  BACKS   OF  FLANGES   ON  TIBES. 

295.  It  is  important  that  the  distance  between  the  backs  of  flanges  on  tires  should 
be  suitable  for  the  gauge  of  the  road  on  which  the  engine  has  to  run.  Increasing  this 
distance  will  obviously  decrease  the  thickness  of  the  flanges,  which  is  not  desirable. 
Decreasing  the  distance  between  the  backs  of  flanges  beyond  a  certain  limit  will 
interfere  with  the  guard  rail  on  the  different  lines  of  railroads.  For  a  gauge  of  4  feet 
8J  inches  the  distance  between  the  backs  of  flanges  is  generally  made  4  feet  5g  inches. 
On  a  few  roads  we  find  this  distance  decreased  to  4  feet  5J  inches ;  and  on  some  other 
roads  increased  to  4  feet  5J  inches. 


CHAPTER  VII. 

MAIN-RODS.— SIDE-RODS.— CRANK-PINS. 
MAIN-RODS  AND   SIDE-RODS. 

296.  Figs.  413,  414  represent  a  main-rod,  and  Figs.  415,  416  represent  a  side-rod. 
Both  were  designed  for  a  four-wheeled  connected  engine,  such  as  shown  in  Fig.  1, 
cylinders  17  x  24  inches. 

The  office  of  the  main-rod  is  to  transmit  motion  from  the  crosshead  to  the  main 
drivers ;  it  connects  the  crosshead  pin  and  the  main  crank-pin,  and  therefore  is  somo- 
tinies  called  the  connecting-rod.  The  term  main-rod  is  mostly  used  in  this  country, 
and  will  be  adopted  in  the  following  descriptions. 

The  office  of  the  side-rod  is  to  transmit  motion  from  the  main  driving  wheels  to 
the  other  driving  wheels ;  it  forms  a  connection  or  coupling  between  the  drivers,  and 
therefore  is  sometimes  called  a  coupling-rod ;  or  again,  because  the  side-rod  on  one 
side  of  the  engine  is  always  parallel  to  the  side-rod  on  the  opposite  side,  it  is  some- 
times called  the  parallel-rod.  But  the  term  side-rod  is  by  far  the  most  popular  one, 
and  will  be  adopted  in  these  articles. 

Several  different  designs  of  main-  and  side-rods  are  in  use,  as  will  soon  be  shown ; 
but  the  designs  here  represented,  we  believe,  are  the  most  common  ones  for  the  class 
of  engine  named.  The  end  of  the  main-rod  marked  F,  Fig.  413,  is  usually  called  the 
front  end ;  and  the  opposite  end,  marked  i?,  the  rear  end. 

The  design  of  the  front  end  of  a  main-rod  depends  on  the  kind  of  crosshead  to  be 
employed.  The  design  of  the  front  end  F  shown  in  Fig.  413  is  only  suitable  for  the 
class  of  crossheads  shown  in  Figs.  234  and  237. 

297.  In  both  rods  the  number  of  bolts  e  e  e  required,  and  their  diameters,  for 
holding  firmly  the  straps  a  a  to  the  butt  ends  of  the  rods,  depend  on  the  magnitude  of 
the  forces  which  these  rods  have   to  transmit.     In  side-rods  we  generally  find  two 
bolts  through  each  strap  to  be  sufficient ;  but  in  main-rods  sometimes  more  than  two 
bolts  through  each  strap  will  be  required.     The  manner  of  determining  the  number  of 
bolts  in  the  main-rod  and  the  diameters  of  bolts  through  all  rods  will  be  explained 
hereafter.     These  bolts  are  generally  turned  to  a  taper  of  J  inch  in  12  inches ;  that  is 
to  say,  the  difference  of  the  two  diameters  12  indies  apart  is  J  of  an  inch;  sometimes 
this  taper  is  only  V,,  inch  in  12  inches.     These  bolts  must  have  an  exti-emely  good 
fit  in,  and  be  driven  tight  into  both  the  straps  and  butt  ends  of  the  rods;    the 
object  of  the  taper  is  simply  this,  that  when  it  becomes  necessary  to  refit  the  brasses 


268 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


b  b  b,  the  bolts  can  be  readily  driven  out  without  impairing  or  upsetting  their  ends. 
Another  advantage  of  the  taper  is  that,  should  these  bolts  at  any  time  become 
somewhat  loose,  they  can  to  some  extent  be  quickly  tightened  by  turning — a  small 
amount — off  the  under  side  of  the  head.  Each  one  of  these  bolts  should  have  two 


MOHEliX   L0( -OMO Tll'K    < 'O.Y.N Tit I'C na\. 


269 


nuts,  mid  since  these  nuts  are  always  placed  at  the  under  side  of  the  strap,  greater 
security  should  be  provided  by  inserting  split  pins  at  the  ends  of  the  bolts. 

_!tS.  The  keys  are  nearly  always  made  of  steel.  Main-rods  have  generally  one 
key  at  each  end ;  in  the  design  shown  in  Fig.  413,  the  key  is  placed  at  the  back  of  the 
crosshead  pin,  and  the  key  in  the  opposite  end  of  the  rod  is  placed  at  the  back  of  the 
main  crank-pin.  This  is  good  practice,  as,  with  this  arrangement  of  keys,  the  brasses 


•• 


/ 

\ 

-  —  ,**  —  t* 

*"/ 

E8-* 

Lu 

fe — tPPHl^t- — >j 

l^-vwjft— ^) 
T*        I 


rfM 

rl    i 


li  "f"  JT'TT  71  "«•'¥-•' 
\^\^l^ 


— fff-P E 


270  MODERN  LOCOMOTirp;   CONSTRUCTION. 

can  be  adjusted  to  some  extent,  without  perceptibly  altering  the  length  of  the  main- 
rod.  Yet  this  arrangement  of  keys  is  by  no  means  a  universal  practice,  as  there  are 
many  main-rods  with  the  key  at  the  front  of  the  crank-pin,  and  one  at  the  back  of 
the  crosshead  pin.  With  this  arrangement  the  rod  will  be  lengthened  when  the  keys 
are  driven  further  into  the  straps. 

Side-rods,  designed  for  four-wheeled  connected  passenger  engines,  generally  have 
two  keys  through  one  end,  and  one  key  through  the  opposite  end ;  and  indeed,  for 
this  class  of  engines  this  arrangement  of  keys  gives  satisfactory  results.  But  in  other 
classes  of  engines  we  sometimes  find  side-rods  with  two  keys  at  each  end,  as  will 
presently  be  shown. 

The  taper  of  the  keys  in  all  rods  varies  from  g  to  1J  inches  in  12  inches.  In  many 
engines  the  small  end  of  the  key  is  threaded,  passed  through  the  guard  d  d,  Fig.  415, 
and  by  means  of  the  nuts  on  each  side  of  the  guard  prevented  from  slipping  out  of 
position.  In  other  engines  we  find  the  keys  held  in  position  by  small  set  screws 
tapped  into  the  straps  or  the  rod,  as  the  case  may  require,  and  for  greater  security,  so 
as  to  prevent  the  key  from  flying  out  should  the  set  screw  at  any  time  become  loose. 
a  split  pin  is  inserted  at  the  small  end  of  the  key. 

Liners  are  inserted  between  the  keys  and  the  brasses.  The  object  of  these  liners 
is  to  prevent  the  keys  from  indenting  or  cutting  the  soft  surface  of  the  brasses. 
These  liners  are  generally  made  of  steel,  sometimes  of  wrought-iron.  Their  thickness 
varies  from  \  to  f  inch. 

299.  Figs.  417,  418  represent  a  main-rod,  and  Figs.  419,  420  represent  a  side-rod 
for  a  four-wheeled  connected  engine,  such  as  is  shown  in  Fig.  1 ;  cylinders  18  x  24 
inches.     The  design  of  the  main-rod  is  similar  to  that  shown  in  Figs.  413,  414,  with  the 
exception  that  the  keys  (Fig.  417)  are.  plain  and  are  held  in  position  by  set  screws. 

The  design  of  the  side-rod,  Figs.  419,  420,  differs  greatly  from  that  shown  in  Fig. 
415.  As  will  be  noticed,  straps  are  not  used  for  this  side-rod.  These  rods  are  forged 
in  one  piece,  the  ends  bored  out,  and  brass  bushings  pressed  in.  The  cross-section  of 
the  rod  here  represented  also  differs  greatly  from  that  of  the  ordinary  side-rod.  The 
cross-section  of  an  ordinary  side-rod  is  rectangular,  similar  to  that  shown  at  A,  Fig 
417,  whereas  the  cross-section  of  the  rod,  Fig.  419,  has  the  appearance  of  an  I.  There 
is  another  distinct  type  illustrated  in  Art.  301,  thus  giving  three  distinct  types  of 
side-rods  which  at  present  are  in  use.  The  advantages  claimed  by  their  respective 
advocates  will  be  considered  later. 

300.  Figs.  421  and  422  represent  a  main-rod  strap  with  key.     This  form  of  strap 
is  very  often  used ;  in  fact,  in  the  majority  of  locomotives  we  find  the  straps  on  all 
the  rods,  excepting  those  on  the  front  end  of  the  main-rod,  made  to  a  rectangular 
form,  as  shown  in  our  illustration.     As  to  the  strap  on  the  front  end  of  the  main-rod, 
we  are  often  compelled  to  round  off  its  end  as  shown  in  Fig.  417,  so  as  to  give  the 
strap  sufficient  clearance  in  the  crosshead  while  the  rod  oscillates  on  the  crosshead 
pin. 

The  length  c  d  of  the  space,  or  that  part  of  the  opening  of  the  strap  marked  A, 
should  be  sufficiently  great  to  admit  the  brasses  and  the  liner  /  as  shown.  The  length 
d  e  of  that  part  of  the  opening  marked  7?  is  determined  by  the  number  of  bolts  through 
the  strap  and  the  width  of  the  key,  as  will  be  presently  explained. 


LOCOMOTIVE   CONSTRUCTION. 


271 


We  occasionally  meet  with  straps  whose  openings  are  of  equal  width  throughout ; 
that  is  to  say,  the  distance  between  t  and  t.2  is  equal  to  that  between  s  and  s2.  But  we 
believe  it  is  safe  to  say  that,  in  a  large  majority  of  straps,  the  thickness  of  the  metal  at 
/  is  greater  than  the  thickness  at  g ;  thus  making  the  width  s  s2  less  than  the  width 
of  the  opening  at  t  t2.  The  straps  are 
planed  so  as  to  make  the  surface  t  paral- 
lel to  t.2 ;  and  s  parallel  to  S2,  leaving  a 
projection  in  line  with  d.  Since  the  por- 
tion of  the  strap  marked  g  is  weakened 
by  the  bolt  holes  i  t,  we  may  be  led  to 
the  conclusion  that  this  strap  is  badly 
proportioned,  and  such  a  conclusion  will 
be  correct,  if  only  the  strength  of  the 
strap  is  considered.  Indeed,  if  we  are 
required  to  design  a  strap  which  shall 
be  simply  of  equal  strength  through- 
out, we  would  make  the  thickness  at  g 
greater  than  that  at/  so  as  to  allow  for 


.*-> 

<K 
I 

HEU 

s 
A 

(I 

e 
B 

Fig.  421 

*t 

tf 

£ 

*l       1 

u 

the  loss  of  strength  caused  by  the  holes 
/  /.  But  in  designing  straps  for  locomo- 
tive rods,  other  considerations  must  be 
taken  into  account.  One  of  the  aims  of  a  locomotive  designer  is  to  design  an  engine 
which  can  be  kept  on  the  road  in  good  working  order  at  a  rnininmm  cost,  and  also  reduce 
as  much  as  possible  the  time  during  which  the  engine  is  kept  out  of  service  for  the  nec- 
essary repairs ;  and  these  are  the  conditions  kept  in  view,  or  which  should  be  kept  in 
view,  when  a  strap  for  a  locomotive  rod  is  to  be  designed.  Consequently,  the  thickness 
of  the  straps  at /is  greater  than  at  g  for  the  following  reason :  In  practice  it  is  found 
that  the  inner  surfaces  *  >s2  will  wear  unevenly,  and  the  brasses  become  loose  long  before 
the  strap  needs  any  other  repairs.  Now  this  extra  thickness  of  metal  at /will  allow 
the  surfaces  s  s.,  to  be  trued  up  or  re-planed  without  touching  the  inner  surfaces  1 12. 
New  brasses  can  then  be  fitted  in  the  space  A  and  the  rod  restored  to  good  working 
order  in  a  comparatively  short  time,  and  at  a  minimum  cost.  On  the  other  hand,  if 
there  had  not  been  any  extra  thickness  of  metal  at  /  then  as  soon  as  the  inner 
surfaces  of  the  strap  become  worn,  they  will  have  to  be  re-planed  from  end  to  end, 
making  the  opening  of  the  strap — that  is,  the  width  from  t  to  t.z — too  wide  for  the  end  of 
the  rod ;  and  therefore,  in  this  case,  the  strap  will  have  to  be  heated  and  upset  at  k,  so 
as  to  close  the  opening  to  fit  the  end  of  the  rod.  In  doing  so  the  holes  i  i  will  be 
thrown  out  of  line,  and  otherwise  cause  considerable  expense  and  an  extra  expenditure 
of  time  (which  always  means  delay  in  getting  the  engine  into  service)  before  the  rod 
can  be  brought  into  working  order. 

The  usual  practice  is  to  make  the  thickness  of  the  strap  at/J  of  an  inch  greater 
than  at  <j ;  occasionally  we  find  the  difference  between  these  thicknesses  to  be  only  -fa 
of  an  inch.  When  there  is  no  key  through  the  end  k  of  the  strap,  the  thickness  at  /•  is 
generally  i  of  an  inch  greater  than  at  //for  small  engines;  and  from  g  to  *  inch 
greater  at  A-  than  at//  for  large  engines.  When  there  is  a  key  thnnigh  the  end  k  of 


272 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


the  strap,  as  shown  in  Fig.  417,  the  thickness  of  the  metal  at  k  is  determined  by  the 
width  of  the  key.  In  small  engines,  say  engines  with  cylinders  14  inches  in  diameter, 
the  thickness  at  k  is  generally  such  as  to  leave  J  inch  metal  at  a  (Fig.  417),  that  is,  J 

inch  metal  at  the  outside  of  the  wide  end 
of  the  key.  As  the  diameters  of  the  cylin- 
ders are  increased,  the  amount  of  metal  at 
a  is  also  increased  at  a  rate  which  will  give 
about  1  inch  metal  at  a  for  engines  with 
cylinders  20  inches  in  diameter.  These 
dimensions  here  given  are  about  the  aver- 
age of  those  in  common  use. 

The  width  of  the  strap  depends  upon 
the  length  of  the  crank-pin  journal,  or  the 
length  of  the  crosshead  pin.  By  deduct- 
ing from  the  lengths  of  these  journals  the 
thickness  of  the  brass  flanges,  the  width  of 
the  strap  is  at  once  determined.  Hence  in 
designing  straps  for  locomotive  main-  or 
side-rods,  about  the  only  calculations  which 
we  have  to  make  are  those  for  finding  the 
correct  diameters  and  the  number  of  bolts 
through  the  straps,  and  the  correct  thick- 
ness of  the  strap  at  g,  Fig.  421. 

301.  Figs.  423,  424  represent  another 
side-rod  designed  for  a  four-wheeled  con- 
nected engine,  such  as  is  shown  in  Fig.  1 ; 
cylinders  17  x  24  inches.  The  whole  de- 
sign of  this  rod  differs  from  the  designs  of 
the  rods  previously  shown.  Straps  are  not 
used,  but  the  keys  are  retained,  so  that  the 
brasses  can  be  adjusted.  The  rod  itself  is 
made  thick  and  narrow  at  the  ends;  and 
thin  and  wide  at  the  center.  Figs.  425, 426, 
427  represent  the  liner  used  between  the 
keys  and  the  brasses.  The  advantages 
claimed  for  this  design  will  be  considered 
hereafter 

MAIN-ROD  BOLTS. 


302.  A  bolt  may  be  subjected  to  a  sin- 
gle or  a  double  shearing  stress.  Thus :  Let 
A  and  B,  Fig.  428,  represent  two  plates 
which  are  connected  by  a  bolt  C.  Now  assume  that  a  force  is  acting  on  the  plate 
A  in  the  direction  of  arrow  2 ;  and  another  force  acting  on  the  plate  B  in  the  direc- 
tion of  the  arrow  3.  In  this  case  the  bolt  is  subjected  to  a  single  shearing  stress, 


LOCOMOTirK    CONSTRUCTION. 


273 


because,  it'  tlic  forces  are  great  enough,  the  bolt  will  be  severed  in  only  one  place,  and 
the  area  through  which  the  bolt  will  be  severed  is  equal  to  the  cross-sectional  area  of 

tin1  bolt. 

In  Fig.  429  the  conditions  are  changed.  Here  we  have  a  bolt  connecting  a  rod 
and  a  strap.  Assume  now  that  a  force  is  acting  on  A  in  the  direction  of  the  arrow  2, 
and  another  force  acting  on  the  strap  in  the  direction  of  the  arrow  3.  In  this  case  the 
bolt  is  subjected  to  a  double  shearing  stress,  because,  if  the  forces  are  great  enough, 
the  bolt  will  be  severed  in  two  places,  and  the  area  through  which  the  bolt  will 
be  severed  will  be  equal  to  twice  the  cross-sectional  area  of  the  bolt.  From  this 
we  infer  that,  if  the  diameter  of  the  bolt  in  Fig.  428  is  equal  to  the  diameter  of  the 
bolt  in  Fig.  429,  the  force  required  to  shear  the  bolt  in  the  latter  figure  will  be  equal 
to  twice  the  force  required  to  shear  the  bolt  in  the  former  figure.  Or,  conversely, 
when  the  bolts  in  both  figures-are  subjected  to  the  same  shearing  forces,  the  cross- 


it 


Fig.  428 


Fig.  430 


Fig. 


sectional  area  of  the  bolt  in  Fig.  428  must  be  equal  to  twice  the  cross-sectional  area 
of  the  bolt  in  Fig.  429.  From  these  considerations  we  learn  that  the  area  sheared  is 
proportional  to  the  shearing  force  to  which  the.  bolt  is  subjected. 

303.  In  determining  the  diameters  and  the  number  of  bolts  through  the  main-rod 
straps,  we  may  assume,  without  impairing  the  results  of  our  calculations,  that  these 
bolts  are  simply  subjected  to  a  double  shearing  stress. 

Here  then  the  question  arises :  How  great  will  be  the  shearing  force  to  which  the 
bolts  through  the  main-rod  straps  are  subjected?  To  this  we  answer :  The  shearing 
force  to  which  the  bolts  are  subjected  will  be  equal  to  the  maximum  pressure  on  the 
main-rod.  To  find  this  maximum  pressure  we  adopt  the  following  graphical  method : 

Fig.  430.  Draw  the  straight  line  a  b,  and  let  it  represent  the  center  line  of  motion 
of  the  piston.  On  this  line  lay  off  any  point  a,  and  let  this  point  represent  the  center 
of  the  driving  axle.  Through  the  point  a  draw  the  line  c/ perpendicular  to  the  line 
a  b,  and  let  the  line  c/represeut  the  vertical  direction  through  which  the  axle  box  can 
move  in  the  pedestal.  On  the  line  c/lay  off  a  point  rf;  this  point  is  to  represent  the 
center  of  the  axle  when  the  axle  box  touches  the  pedestal  cap,  consequently  the 
distance  between  the  points  a  and  <l  must  be  equal  to  the  distance  between  the  center 
line  of  motion  a  6,  and  the  lowest  position  that  the  center  uf  the  axle  can  occupy  in  the 
pedestal.  Again,  on  the  line  cf  lay  off  a  point  <-;  this  point  is  to  represent  the  center 


274  MODERN  LOCOMOTITE   CONSTRUCTION. 

of  the  axle  when  the  axle  box  touches  the  upper  end  of  the  pedestal,  consequently  the 
distance  between  the  points  a  and  e  must  be  equal  to  the  distance  between  the  center 
line  of  motion  a  b  and  the  highest  position  that  the  center  of  the  axle  can  occupy  in 
the  pedestal.  If  now  we  find  that  the  distance  a  d  is  greater  than  the  distance  a  e, 
then  from  the  point  d  lay  off  on  the  line  c/a  point  c;  the  distance  between  the  points 
d  and  c  must  be  equal  to  one-half  the  stroke,  or,  in  other  words,  equal  to  the  length  of 
the  crank.  If,  on  the  other  hand,  the  distance  a  e  is  greater  than  the  distance  a  d, 
then  from  the  point  e  lay  off  on  the  line  c/a  point/;  the  distance  between  the  points 
e  and  /  must  be  equal  to  one-half  the  stroke.  But  generally  in  locomotives,  the 
distance  a  d  will  be  found  to  be  greater  than  a  e,  hence  we  shall  confine  our  attention 
to  that  which  happens  below  the  center  line  of  motion  a  I.  From  the  point  c  as  a 
center,  and  with  a  radius  equal  to  the  length  of  the  connecting-rod,  describe  an  arc 
cutting  the  line  a  b  in  the  point  I ;  join  the  points  c  and  b  by  a  straight  line,  and  thus 
completing  the  right-angled  triangle  a  c  I.  Now  the  length  of  the  side  a  b  of  this 
triangle  will  represent  the  total  steam  pressure  on  the  piston ;  the  length  of  the  side 
c  b  will  represent  the  maximum  pressure  on  the  main-rod,  and  according  to  what  has 
been  stated  before,  the  length  of  the  side  c  b  will  also  represent  the  shearing  force  to 
which  the  bolts  in  the  rod  are  subjected.  Therefore,  as  soon  as  we  know  the  press- 
ure which  the  length  of  the  side  a  b  does  represent,  we  will  have  no  difficulty  in 
determining  the  shearing  force  on  the  bolts ;  because,  as  we  have  seen,  the  line  a  b 
I'epresents  the  total  steam  pressure,  and  this  pressure  is  easily  found  by  multiplying 
the  area  of  the  piston  by  the  pressure  per  square  inch ;  the  pressure  on  the  main- 
rod  due  to  the  steam  pressure  will  be  as  much  greater  than  the  total  pressure  on  the 
piston,  as  the  length  of  the  line  c  b  is  greater  than  the  length  of  the  line  a  b,  and 
consequently  this  pressiire  can  be  found  by  a  simple  rule  of  proportion.  In  order 
to  show  plainly  the  manner  of  applying  these  principles,  we  will  take  the  following 
example : 

EXAMPLE  84. — It  is  required  to  find  the  shearing  force  to  which  the  main-rod 
bolts  are  subjected  in  a  locomotive  whose  cylinders  are  16  inches  diameter;  stroke, 
24  inches ;  maximum  steam  pressure  in  the  cylinders  is  120  pounds  per  square 
inch ;  length  of  the  connecting-rod,  84  inches ;  the  distance  (a  d,  Fig.  430)  below 
the  center  line  of  motion  of  the  piston  through  which  the  center  of  axle  can  move  is 
3  inches. 

In  the  first  place,  let  us  find  the  lengths  of  all  the  sides  of  a  right-angled  triangle, 
such  as  shown  in  Fig.  430,  whose  hypothenuse  shall  represent  the  pressure  on  the 
main-rod  due  to  the  steam  pressure  given  in  our  example. 

The  lengths  of  two  sides  of  this  triangle  are  already  known,  for  we  know  that  the 
side,  or  hypothenuse,  c  b  must  be  equal  to  the  length  of  the  connecting-rod,  namely, 
84  inches.  The  length  of  the  side  a  c  must  be  equal  to  the  sum  of  the  distance  a  d, 
which  is  3  inches,  and  one-half  the  stroke,  which  is  12  inches ;  hence,  the  side  a  c  will 
be  12  +  3  =  15  inches.  The  length  of  the  side  a  b  we  must  find  by  the  well-known 
rule  given  in  geometry  for  finding  any  side  of  a  right-angled  triangle  when  two  of 
its  sides  are  known.  In  the  case  before  us  we  must  subtract  the  square  of  the  side 
a  c  from  the  square  of  the  side  c  b,  and  extract  the  square  root  of  the  remainder. 

The  square  of  the  side  b  c  is  equal  to  84  x  84  =  7056. 


MODERN  LOCOMOTIVE  CONSTRUCTION.  275 

The  square  of  the  side  a  c  is  equal  to  15  x  15  =  225. 

Subtracting  the  latter  square  from  the  former,  we  have  7056  —  225  =  6831. 

Extracting  the  square  root  of  6,831,  we  have 

\/6831  =  82.64  inches, 

which  is  the  length  of  the  side  a  b. 

The  area  of  a  piston  16  inches  in  diameter  is  equal  to  201.06  square  inches,  and 
the  total  steam  pressure  on  one  piston  will  be  equal  to  201.06  x  120  =  24127.20 
pounds. 

Hence  line  a  b,  which  we  have  found  to  be  82.64  inches  long,  represents  24,127.20 
pounds.  The  line  c  b  we  know  to  be  84  inches  long,  and  since  the  pressure  represented 
by  the  side  c  b  will  be  as  much  greater  than  24,127.20  pounds  as  c  b  is  longer  than  a  b, 

we  have 

84  x  24127.20 

8264       -  =  24524.25  pounds, 

which  is  the  pressure  represented  by  the  line  c  b,  and  consequently  is  the  press- 
ure on  the  main-rod,  or  the  shearing  force  to  which  the  bolts  in  the  main-rod  are 
subjected. 

304.  Our  next  step  will  be  to  find  the  area  necessary  to  resist  this  shearing  force, 
and  this  area  will  be  equal  to  twice  the  total  cross-sectional  area  of  all  the  bolts 
through  one  end  of  the  rod.  In  order  to  find  this  area  we  must  first  establish  the 
stress  which  should  be  allowed  per  square  inch,  and  this  can  best  be  established  by 
finding  the  stress  allowed  by  builders.  For  the  sake  of  simplicity,  we  shall  consider 
that  the  bolts  are  subjected  to  a  shearing  stress  only,  and  shall  consider  that  the 
total  stress  is  equal  to  the  total  pressure  on  the  main-rod.  Now,  assuming,  as  we  have 
done  before,  that  the  pressure  in  the  cylinders  on  locomotives  at  present  in  service, 
whose  safety  valves  are  set  to  130  pounds,  will  be  120  pounds  per  square  inch,  and 
then  determining  by  calculation  the  stress  (due  to  the  pressure  of  120  pounds)  which 
has  been  allowed  per  square  inch  of  cross-sectional  area  of  the  main-rod  bolts,  we  find 
that  it  varies  in  different  locomotives  from  7,000  to  9,000  pounds,  and  in  one  case  we 
found  it  to  be  as  high  as  12,000  pounds  per  square  inch.  Our  experience  leads  us  to 
believe  that,  when  the  best  quality  of  iron  is  used,  8,000  pounds  per  square  inch  of  the 
cross-sectional  area  of  the  main-rod  bolts  will  give  good  results,  and  should  be  adopted. 
The  best  quality  of  wrought-iron  of  which  these  bolts  are  generally  made  possesses  an 
ultimate  shearing  strength  of  about  56,000  pounds  per  square  inch ;  if  we  now  assume 
that  these  bolts  are  simply  subjected  to  a  shearing  stress,  and  allow  8,000  pounds  per 
square  inch  of  section,  we  adopt  7  as  a  factor  of  safety.  Hence,  if  the  ultimate  shear- 
ing strength  of  the  iron  is  less  than  56,000  pounds  per  square  inch,  say  it  is  only 
49,000  pounds,  then  the  safe  working  stress  to  be  allowed  per  square  inch  will  be 
equal  to 

49000 

=  <000  pounds. 

From  the  foregoing  remarks  we  conclude  that  twice  the  total  cross-sectional  area 
of  the  bolts — that  is,  the  area  which  has  to  resist  the  shearing  force — must  be  proper- 


276  MODERN  LOCOMOTIVE   CONSTRUCTION. 

tioned  so  that  the  stress  per  square  inch  will  be  8,000  pounds,  provided  the  best  quality 
of  iron  is  used. 

After  this,  when  the  number  and  diameters  of  the  bolts  are  to  be  established,  we 
may  be  compelled  to  change  this  stress  of  8,000  pounds  per  square  inch,  because  the 
general  practice  in  locomotive  building  is  to  avoid  -fa  and  ^  of  an  inch  in  the  diame- 
ters of  the  bolts,  and  therefore,  in  establishing  the  nearest  number  of  bolts  and 
adopting  the  nearest  practical  diameters  to  those  for  which  the  result  of  the  calculations 
call,  the  stress  may  be  somewhat  greater  or  less  than  8,000  pounds  per  square  inch  ; 
indeed,  we  may  have  to  be  satisfied  as  long  as  the  stress  falls  within  the  limits  of 
7,500  to  8,500  per  square  inch  for  the  best  quality  of  iron.  For  an  inferior  quality  of 
iron  the  stress  should  be  less. 

Now,  having  established  the  stress  to  be  allowed  per  square  inch,  we  can  easily 
find  the  total  cross-sectional  area  of  the  main  -rod  bolts  through  one  end  of  the  rod  by 
the  following  rule  : 

RULE  54.  —  Divide  the  total  pressure  on  the  main-rod,  as  found  by  the  diagram, 
Fig.  430,  by  8,000  ;  the  quotient  will  be  twice  the  total  cross-sectional  area  of  the  main- 
rod  bolts  through  one  of  its  ends. 

EXAMPLE  85.  —  The  total  pressure  on  the  main-rod  given  in  Example  84  is  24,524.25 
pounds  ;  what  will  be  twice  the  total  cross-sectional  area  of  the  bolts  t 

24524.25 

=  3.065  square  inches. 


305.  These  bolts,  as  we  have  stated  befoi*e,  are  tapered,  hence  when  we  speak  of 
the  diameters  of  the  bolts  we  mean  their  small  diameters.  In  a  large  number  of 
locomotives  the  rear  strap  is  wider  than  the  front  or  crosshead  strap.  Careful 
observation  seems  to  indicate  that  the  diameter  of  any  one  of  these  bolts  should  be 
equal  to  about  one-third  the  width  of  the  rear  strap  ;  in  some  engines  the  diameters 
are  a  little  greater,  and  in  others  a  little  less,  than  this  proportion.  We  may  therefore 
assume  that  it  is  good  practice  to  make  the  diameter  of  the  bolt  (as  near  as  possible 
consistent  with  the  avoidance  of  -fa  or  -3\  of  an  inch  in  the  diameter)  equal  to  one- 
third  of  the  width  of  the  rear  strap.  Hence  we  have  the  following  rule  : 

RULE  55.  —  Divide  the  width  of  the  rear  strap  in  inches  by  3  ;  the  quotient  will  be 
the  diameter  of  the  bolt. 

EXAMPLE  86.  —  The  width  of  the  strap  is  2f  inches  ;  find  the  diameter  of  the  bolts. 


which  is  nearly  equal  to  ff  inch.  Avoiding  -fa  or  -fa  inch,  we  say  the  diameter  of  the 
bolts  should  be  y§-  inch.  Some  builders  will  make  these  bolts  |  inch  diameter. 

306.  When  we  know  the  diameters  of  the  bolts,  and  also  the  total  cross-sectional 
area  which  has  to  resist  the  shearing  force,  the  number  is  easily  determined  by  the 
following  rule  : 

RULE  56.  —  Divide  the  total  cross-sectional  area  of  the  bolts,  as  found  by  Rule  54, 
by  twice  the  cross-sectional  area  of  one  bolt  ;  the  quotient  will  be  the  number  of  bolts 
required. 


MODERN  LOCOMOTIVE   CONSTRUCTION.  277 

EXAMPLE  87.  —  The  total  cross-sectional  area  of  the  bolts  subjected  to  shearing 
stivss  is  3.065  square  indies,  the  diameter  of  each  bolt  is  if  inch.  How  many  bolts 
will  be  required  I  The  cross-sectional  area  of  a  if  -inch  bolt  is  .69  square  inch  ;  hence 

3.065 

=  2'2'  sa?  2  bolts' 


In  this  case  the  stress  per  inch,  when  the  taper  of  the  bolt  is  taken  into  account, 
which,  for  the  sake  of  simplicity,  has  been  left  out  of  consideration,  will  not  exceed 
8,500  pounds. 

EXAMPLE  88.  —  The  total  pressure  on  one  main-rod  of  a  locomotive  —  cylinders 
18  x  24  —  is  31,000  pounds  ;  the  width  of  the  strap  is  2$  inches  ;  find  the  diameter  and 
number  of  bolts. 

According  to  Rule  54,  twice  the  total  cross-sectional  area  of  the  bolts  should  be 

31000 

'•=  3.87  square  inches. 


According  to  Rule  55,  the  diameter  of  each  bolt  should  be 

°  75 

-y-  =  .916,  say  f|  inch. 

The  cross-sectional  area  of  a  bolt  if  inch  diameter  is  .69  square  inch;  hence, 
according  to  Rule  56,  the  number  of  bolts  required  will  be 

3  87 

2~^-gg  =  2-8>  say  3  bolts- 

Some  locomotive  builders  use  three  bolts  &  inch  in  diameter  for  the  same  size  of 
engine. 

EXAMPLE  89.  —  The  total  pressure  on  one  main-rod  of  a  locomotive  —  cylinders 
20  x  24  —  is  38,000  poimds,  strap  3  inches  wide;  find  the  diameter  and  number  of 
bolts. 

According  to  Rule  54,  the  total  cross-sectional  area  of  the  bolts  will  be 

38000 


8000 


=  4.75. 


According  to  Rule  55,  the  diameter  of  the  bolts  will  be  1  inch.  The  cross-sectional 
area  of  a  bolt  1  inch  in  diameter  is  .7854;  hence,  according  to  Rule  56,  the  number  of 

bolts  through  each  end  of  imod  will  be 

4.75 

2  x  .7854  ~ 

There  are  engines  of  the  same  size  running  with  only  2  bolts  1  inch  in  diameter 
through  each  end.  We  believe  that  3  bolts,  1  inch  diameter,  for  a  maximum  steam 
pressure  of  120  pounds  in  a  cylinder  20  inches  diameter,  will  be  safer  and  give  better 
results. 

EXAMPLE  90. — What  should  be  the  diameter  of  the  bolt  shown  in  the  rear  end  of 
the  main-rod,  Figs.  431,  432!  The  cylinder  is  12  x  16  inches;  maximum  steam 
pressure  in  the  cylinder,  120  pounds;  the  distance  through  which  the  axle  box  can 
move  below  the  center  line  of  motion  of  the  piston  is  2  inches. 


278  MODERN  LOCOMOTirE   CONSTRUCTION. 

Although  the  design  of  this  rod  is  different  from  any  we  have  previously  shown, 
the  foregoing  rules  are  applicable.  Here  we  will  first  have  to  find  the  total  maximum 
pressure  on  the  rod,  as  explained  in  Art.  303,  and  illustrated  in  diagram,  Fig.  430. 
The  area  of  a  piston  12  inches  in  diameter  is  113.1  square  inches,  hence  the  total 
pressure  on  the  piston  is  equal  to 

113.1  x  120  =  13572  pounds ; 

consequently  the  line  a  b  in  our  diagram,  Fig.  430,  represents  13,572  pounds.  But  the 
length  of  the  line  a  I  is  not  yet  known,  and  must  be  found  according  to  the  instruc- 
tions given  in  Art.  303. 

In  order  to  suit  the  conditions  given  in  the  example,  the  length  of  the  line  a  c 
must  be  10  inches — that  is,  the  sum  of  half  the  stroke  (8  inches)  and  the  distance 
(2  inches)  through  which  the  axle  box  can  move  below  the  center  line  of  motion 
of  the  piston.  The  length  of  the  line  c  b  is  equal  to  the  length  of  the  connecting-rod 
— that  is,  the  distance  between  the  centers  of  journals — and,  as  will  be  seen  in  Figs. 
431,  432,  this  distance,  or  length  of  the  connecting-rod,  is  60  inches.  Therefore  the 
length  of  the  line  a  I  will  be  equal  to 


V602  -  10*  =  59.16  inches. 
Consequently,  according  to  Art.  303,  the  line  c  b  will  represent 

60  x  13572 

^n  1g —  =  13^64./  pounds, 
oy.lb 

which  is  the  maximum  pressure  on  the  main-rod. 

Twice  the  total  cross-sectional  area  of  the  bolt  to  resist   shearing  must  be, 

according  to  Rule  54. 

13764.7 
QAAA    =  1.72  square  inches, 

oUUU 

and 

1.72 

—^~  =  .86  square  inch, 

which  will  be  the  cross-sectional  area  of  the  bolt.  The  diameter  of  the  bolt  whose 
cross-sectional  area  is  .86  square  inch  is  nearly  l-j^  inches,  agreeing  very  closely  with 
diameter  of  the  bolt  given  in  the  illustration. 

307.  Figs.  431,  432  represent  a  main-rod,  and  Figs.  433,  434  a  side-rod  used  on 
engines  running  on  one  of  the  elevated  roads ;  the  cylinders  of  these  engines  are  12 
inches  diameter  and  16  inches  stroke.  The  rods  are  very  good  ones,  and  well  adapted 
for  these  engines.  The  design  of  the  main-rod  differs  from  the  designs  of  rods  previ- 
ously given.  Straps  are  not  used;  the  rear  end  is  an  open  one,  which,  after  the 
bearings  have  been  placed  in  position,  is  closed  by  the  steel  block  A,  firmly  held  in 
position  by  the  bolt  B.  This  bolt  must  be  strong  enough  to  resist  the  pull  of  the  rod. 
The  front  end  of  the  rod  is  designed  to  suit  a  class  of  crossheads  with  a  single  slide, 
such  as  is  shown  in  Fig.  248. 

The  design  of  the  side-rod  also  differs  from  the  designs  of  side-rods  previously 
given.  Here  the  straps  have  been  retained,  but  keys  are  not  used. 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


279 


308.  In  the  previous  examples  we  have  made  the  diameters  of  the  maiii-rod  bolts 
i -q  i  ml  to  oiie-third  of  the  width  of  the  rear  strap  as  nearly  as  possible.  The  rear  strap, 
as  we  have  stated  before,  is  generally  made  a  little  wider  than  the  front  one,  so  as  to 
make  the  former  stronger  than  the  latter;  the  reason  for  this  will  be  presently 


m  rr>  •----$••• 

i^jTr 

<H_ 


i  I 

-_^V-j,. 

TV  Kt 


MM 


Fig.  433 


- 

J^p       =5*^=1 

I       '       1  II 


& 


i         I  .- 

* nef-.~. <| 


Fig.  434 


--XM — 


i  i 

1 

!j£  Searings  of 

•^ 7F — \-fhoiphor  Bronze     ,, 

'&  \     Taper  of  Bolts  X  In  19 


N~'    I              f        J  •'' 
^-4 «j<f ix^ 

1  M 


Steel 


1  „  i  sa 


. 

-331  _-_  71 

H~T  / 


r.  432 


i  « 
j,. jg* 


1 


explained.  But,  although  the  width  of  the  rear  strap  differs  from  the  width  of 
the  front  one,  the  diameters  of  tin-  bolts  are  nearly  always  m.-ulf  Hie  same  for 
each  ond. 

Again,  iii  the  previous  examples,  the  diameters  of  the  bolts  and  their  number 


280  MODERN  LOCOMOTIVE   CONSTRUCTION. 

were  made  to  suit  a  maximum  steam  pressure  of  120  pounds  per  square  inch  of  piston. 
The  stress  of  8,000  pounds  per  square  inch  of  the  cross-sectional  area  of  the  bolts  was 
also  established ;  and  since  this  stress  should  not  be  changed,  no  matter  how  great  or 
small  the  maximum  steam  pressure  in  the  cylinders  may  be,  it  follows  that  the  rules 
previously  given  for  determining  the  diameters  and  number  of  bolts  through  the 
main-rod  are  also  applicable  when  the  maximum  steam  pressure  in  the  cylinder  is 
greater  or  less  than  120  pounds  per  square  inch.  When  the  steam  pressure  is  greater 
than  120  pounds  per  square  inch,  then  we  shall  generally  require  three  bolts  through 
each  end  of  the  main-rod  for  locomotives  having  cylinders  17  inches  in  diameter  and 
upwards.  Since  the  greatest  number  of  bolts  used  through  each  strap  is  three  (at  least 
we  have  not  met  with  cases  in  which  this  number  was  exceeded),  the  problems  for 
solution  may  present  themselves  in  the  following  form. 

EXAMPLE  91. — Three  bolts  are  to  be  used  through  one  end  of  a  main-rod  for  a 
locomotive  having  cylinders  18  x  24  inches.  The  length  of  the  main-rod  is  90  inches ; 
the  maximum  steam  pressure  in  the  cylinder  is  160  pounds;  the  vertical  distance 
below  the  center  line  of  motion  of  the  piston  through  which  the  center  of  axle  can 
move  is  3  inches.  It  is  required  to  find  the  diameter  of  the  bolts. 

In  Art.  303  we  have  stated  that  the  shearing  force  to  which  the  bolts  through  one 
end  of  the  main-rod  are  subjected  is  represented  by  the  length  of  the  hypothenuse  c  b 
of  a  right-angled  triangle,  Fig.  430,  in  which  the  length  of  the  side  a  b  represents  the 
total  steam  pressure  on  the  piston ;  hence  our  first  step,  in  the  solution  of  our  problem, 
will  be  to  find  the  lengths  of  the  sides  of  such  a  triangle.  In  Art.  303  we  also  have 
stated  that  the  length  of  the  hypothenuse  must  be  equal  to  the  length  of  the  connect- 
ing-rod ;  hence,  in  this  case  the  length  of  the  hypothenuse  must  be  90  inches.  The 
length  of  the  side  a  c,  that  is,  the  shortest  side,  must  be  equal  to  the  sum  of  the  length 
of  the  crank,  and  the  vertical  distance  below  the  center  line  of  motion  through  which 
the  center  of  the  axle  can  move ;  hence,  according  to  the  conditions  given  in  the 
example,  the  length  of  the  side  a  c  must  be  equal  to  12  +  3  =  15  inches.  The  length 
of  the  side  a  b  must  be  found  by  calculation  as  explained  in  Art.  303,  thus : 

Eemembering  that  the  length  of  the  hypothenuse  is  90  inches,  and  the  length  of 
the  side  a  c  is  15  inches,  we  have 


V902    -  152  =  88.74  inches, 

which  is  the  length  of  the  side  a  b. 

Now,  having  found  the  lengths  of  the  sides  of  the  right-angled  triangle,  we  can 
easily  determine  the  shearing  force  to  which  the  bolts  are  subjected ;  thus : 

The  area  of  a  piston  18  inches  in  diameter  is  254.46  square  inches ;  therefore  the 
total  pressure  on  the  piston  at  160  pounds  per  square  inch  will  be 

254.46  x  160  =  40713.60  pounds. 

According  to  the  remarks  in  Art.  303,  the  length  of  the  side  a  b,  namely,  88.74 
inches,  represents  a  pressure  of  40713.60  pounds ;  that  is  to  say,  the  length  of  the  side 
a  b  represents  the  total  steam  pressure  on  the  piston.  But  we  also  know  that  the 
maximum  pressure  on  the  main-i*od  is  represented  by  the  length  of  the  hypothenuse 
of  the  same  triangle,  and  that  the  maximum  pressure  on  the  main-rod  is  as  much 


MODERN  LOCOMOTIVE   CONSTRUCTION.  281 

greater  as  the  length  of  the  hypothenuse  is  greater  than  the  length  of  the  side  a  b. 

Hence  we  have 

40713.60  x  90 

Og^      -  =  41291.68  pounds, 

which  is  the  maximum  pressiire  on  the  main-rod;  and  this  pressure  also  represents 
the  total  shearing  force  to  which  the  three  bolts  are  subjected. 

Since  the  stress  per  square  inch  is  to  be  8,000  pounds,  twice  the  total  cross- 
sectional  area  of  the  bolts — that  is,  the  whole  area  which  has  to  resist  the  shearing 
force — is  found  by  Eule  54,  and  is  equal  to 


41291.68 

-  =  5.16  square  inches. 

o(JUU 


But  our  example  requires  that  three  bolts  shall  resist  the  shearing  force,  and  since  the 
bolts  are  subjected  to  a  double  shearing  stress,  twice  the  total  area  of  the  bolts  which 
has  to  resist  the  shearing  force  will  be  equal  to  six  times  the  cross-sectional  area  of  one 
bolt ;  hence  we  have 

5'16        Qp 

-~-    =  .86  square  inch, 

which  is  the  cross-sectional  area  of  one  bolt.  The  diameter  of  a  circle  whose  area  is 
.86  inch  is  nearly  l^-  inches;  consequently  the  diameter  of  the  bolts  should  be  liV 
inches. 

309.  Figs.  435,  436  represent  a  main-rod,  and  Figs.  437,  438  represent  a  side-rod 
for  a  Mogul  engine,  that  is  to  say,  a  locomotive  having  six  driving  wheels  and  a  pony, 
or  two-wheeled  truck. 

In  these  engines  the  side-rods  have  to  connect  three  pairs  of  driving  wheels,  and 
transmit  motion  to  the  front  and  rear  pair.  We  have  already  seen  that  the  design  of 
the  pedestal  permits  the  driving  axles  to  move  up  and  down,  so  as  to  enable  the  driving 
wheels  to  adjust  themselves  to  any  unevenness  of  the  track,  consequently  the  centers 
of  the  axles  will  not  always  lie  in  a  horizontal  line.  Under  these  conditions  we  cannot 
connect  the  three  driving  wheels  on  one  side  of  the  engine  by  a  side-rod  in  which  the 
crank-pin  bearings  are  held  rigidly  in  line.  Consequently  the  side-rods  in  Mogul 
engines  are  made  in  two  pieces,  forming  a  front  and  rear  side-rod  for  each  side  of  the 
engine,  as  shown  in  Figs.  437,  438. 

The  front  and  rear  side-rods  are  connected  by  the  pin  S,  which  is  placed  at  the 
back  of  the  main-pin  M.  As  far  as  the  working  of  the  engine  is  concerned,  the  pin  8 
could  be  placed  in  front  of  the  main-pin  M ;  but  for  the  sake  of  convenience  it  is 
always  best  to  place  it  at  the  back  of  the  main  crank-pin,  so  that  it  will  at  no  time  be 
covered  by  the  main-rod,  which,  in  nearly  all  Mogul  engines,  is  placed  outside  of  the 
side-rods.  In  fact,  it  often  happens  that,  with  the  main-rods  placed  outside  of  the 
side-rods,  the  available  space  between  the  side-  and  main-rods  will  not  be  sufficient  for 
the  nuts  on  the  pin  S,  and  consequently  this  pin  must  be  placed  at  the  back  of  the 
main-pin  M  as  shown.  The  pin  S  should  also  be  placed  as  near  as  possible  to  the 
main  crank-pin,  so  as  to  reduce  the  stress  to  which  the  strap  around  the  main-pin  is 
subjected  to  a  minimum. 


282 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


K- — ,-,'Kio >•  vm 

r — i  ,f°'i — T*J* 


j, K-8 


«-^»~*« 


r        »•"• 

^ %ot » 


MODERX  LOCOMOTIVE   CONSTRUCTION.  1>S:; 

From  the  foregoing  remarks  we  may  conclude  that  in  Mogul  engines  the  front 
side-rods  contain  the  bearings  for  the  front  crank-pin  F  and  for  the  main  crank-pin 
M;  the  rear  side-rod  simply  contains  the  bearing  for  the  rear  crank-pin  JR. 

MAIN-ROD   STRAPS. 

310.  In  Art.  300  we  have  stated  that  in  designing  the  straps  for  locomotive  main- 
ami  side-rods,  about  the  only  calculations  which  we  will  have  to  make  are  those  for 
finding  the  correct  diameters  for  the  bolts,  number  required,  and  the  thickness  of  the 
strap  represented  by  </,  in  Fig.  421. 

The  rules  for  finding  the  number  of  bolts  and  their  diameters  have  been  given ; 
hence  it  now  only  remains  for  us  to  establish  rules  for  finding  the  thickness  of  the 
strap  at  g. 

The  weakest  part  of  any  locomotive  main-  or  side-rod  strap  is  in  the  plane  x  y  or 
#2  y.,  (Fig.  422),  passed  through  the  axis  of  the  bolt.  Consequently,  when  the  width  of 
the  sti'ap  is  given,  the  thickness  at  <j  must  be  determined  by  calculation  so  as  to  give  a 
sufficient  cross-sectional  area  in  the  plane  x  y  to  resist  the  forces  acting  on  the  strap. 
Of  course,  in  the  foregoing  remarks  we  have  assumed  that  the  key  has  the  best 
possible  shape  for  the  work  it  has  to  do  (more  of  this  hereafter),  and  that  its  thickness, 
as  well  as  the  diameter  of  the  hole  for  the  oil-cup,  does  not  exceed  the  diameter  of  the 
bolts. 

The  stress  *  in  the  main-rod  straps  is  greater  than  that  in  the  side-rod  straps ;  we 
shall,  therefore,  in  this  article,  consider  the  strength  of  the  main-rod  straps  only. 

Both  the  front  and  the  rear  straps  are  subjected  to  a  tensile  force — that  is,  a  force 
tending  to  tear  the  straps.  But  the  centrifugal  force  to  which  the  rear  end  of  the  rod 
is  subjected  brings  into  play  on  both  straps  a  force  which  acts  in  a  direction  perpen- 
dicular to  the  length  of  the  rod,  or,  in  short,  a  transverse  force ;  this  force  acts  with  the 
greatest  intensity  at  the  back  end  of  the  rod,  and  becomes  zero  at  the  center  of  the 
crosshead  pin.  Yet  there  are  portions  of  the  front  strap,  which  on  account  of  their  dis- 
tances from  the  center  of  the  crosshead  pin,  are  more  or  less  subjected  to  a  transverse 
force.  Also,  noticing  that  the  bolts  through  the  straps  must  necessarily  be  placed  at 
some  distance  from  the  center  of  the  pins,  the  transverse  forces  on  both  straps  will  act 
with  a  leverage,  and  therefore  will  act  with  a  greater  effect,  and  cause  a  greater  stress 
in  the  straps,  than  they  would  do  if  the  bolts  could  be  closer  to  the  center  of  the  pins. 
Consequently,  these  latter  forces  must  not  be  neglected  in  our  calculations. 

But  the  transverse  force  acting  on  the  rear  strap  is  greater  than  the  transverse 
force  acting  on  the  front  one;  and  since  the  tensile  forces  are  equal  on  both  straps,  it 
follows  that  the  total  stress  in  the  rear  strap  is  greater  than  the  total  stress  in  the 
front  one. 

Again,  the  stress  per  square  inch  of  cross-sectional  area  should  be  the  same  in 
both  straps ;  hence,  it  follows  that,  since  the  total  stress  is  greater  in  the  rear  strap  than 
that  in  the  front  one,  the  former  should  be  made  stronger  than  the  latter,  and  this 
we  often  find  to  be  the  case  in  practice.  But  many  master-mechanics  prefer  to 

•"  Stress"  may  be  defined  as  the  resistance  to  the  alteration  of  form,  and  in  this  sense  the  word  "stress"  is 
here  used. 


284  MODERN  LOCOMOTIVE   CONSTRUCTION. 

make  these  straps  of  equal  strength,  for  the  following  reason:  The  crosshead  pin 
•will  wear  in  time  to  an  oval  form,  making  it  smaller  in  the  direction  of  the  axis  of 
the  cylinder.  If  now  the  brasses  in  the  front  end  of  the  rod  do  not  touch  each 
other,  as  is  often  the  case,  and  the  rod  begins  to  pound,  the  key  may  be  driven 
accidentally  or  carelessly  too  tight,  causing  the  brasses  to  bind  on  the  large  part 
of  the  crosshead  pin,  throwing  on  the  front  strap  an  extra  stress,  which  may,  un- 
less the  front  strap  is  strong  enough  to  resist  this  extra  force,  cause  considerable 
damage. 

There  is  another  fact  which  must  not  be  lost  sight  of  in  proportioning  a  strap, 
namely,  the  bolt  holes  in  the  strap  will  in  time  wear  to  an  oval  shape,  and  then  they 
must  be  re-reamed,  thereby  weakening  the  strap ;  therefore  it  is  necessary,  in  design- 
ing a  strap,  to  make  a  slight  allowance  for  re-reaming. 

Now,  in  order  to  determine  accurately  the  thickness  g  of  the  strap,  we  should  know 
the  exact  magnitudes  of  the  forces  to  which  they  are  subjected.  But  to  find  the  exact 
magnitudes  of  these  forces  will  be  a  difficult  matter.  Hence,  taking  all  the  given 
conditions  into  consideration,  it  will  be  an  easier  way,  and  probably  the  most  practical 
one,  to  base  our  calculations  on  the  proportions  of  straps  used  at  present  in  locomo- 
tives running  at  high  speeds,  doing  excellent  service. 

We  have  seen  that  the  weakest  part  of  the  strap  is  in  the  plane  x  y;  hence, 
our  first  step  will  be  to  determine  the  cross-sectional  area  of  the  strap  in  this  plane. 

Suppose,  for  the  sake  of  simplicity,  that  the  strap  is  subjected  to  a  tensile  force — 
that  is  to  say,  a  force  acting  in  the  direction  of  the  length  of  the  rod,  or  in  a  direction 
indicated  by  the  arrow  in  Fig.  421 — tending  to  pull  the  strap  apart.  In  this  strap  we 
have  two  bolts,  which  we  may  assume  resist  aii  equal  amount  of  the  pull,  or,  in  other 
words,  each  bolt  resists  one-half  of  the  pull.  We  therefore  conclude  that  the  metal  of 
the  strap  in  the  plane  x2  y2  has  to  resist  one-half  of  the  tensile  force  or  pull  to  which 
the  strap  is  subjected.  But  the  force,  acting  on  the  metal  in  the  plane  x.2  y-2,  is 
transmitted  to  the  plane  x  y,  and  this  plane  must  also  resist  the  tensile  force  trans- 
mitted to  it  by  its  own  bolt.  Therefore  the  metal  of  the  strap  in  the  plane  x  y  must 
resist  the  whole  tensile  force  to  which  the  strap  is  subjected ;  and  the  area  of  this 
metal  must  be  proportioned  accordingly.  If  we  had  three  bolts  through  the  strap,  as 
is  often  the  case,  the  conditions  would  not  be  changed — that  is  to  say,  the  weakest 
part  of  the  strap,  namely,  its  portion  in  the  plane  x  y,  must  be  made  strong  enough  to 
resist  the  whole  tensile  force,  and  treated  just  the  same  as  if  no  other  portion  of  the 
strap  had  to  resist  any  part  of  the  force  acting  upon  it. 

If,  then,  the  straps  are  subjected  to  a  tensile  force  pure  and  simple,  as  we  have 
supposed,  we  can  easily  find  the  cross-sectional  area  at  x  y  by  allowing  a  stress  of 
10,000  pounds  per  square  inch,  as  we  have  done,  on  the  weakest  part  of  a  piston-rod, 
and  consequently  obtain  the  number  of  square  inches  in  the  cross-sectional  area,  by 
dividing  the  maximum  pressure  on  the  main-rod  by  10,000 ;  the  quotient  will  be  the 
required  area  through  both  wings  y  y.,  of  the  strap. 

But  we  have  seen  that  these  straps  are  subjected  to  forces  acting  in  a  direction 
perpendicular  to  the  length  of  the  rod,  and  these  forces  must  not  be  neglected.  But 
since  it  is  difficult  to  determine  the  exact  magnitude  of  these  forces,  we  make  allow- 
ances for  them  by  reducing  the  stress  per  square  inch  due  to  the  tensile  force  acting 


MODERN  LOCOMOTIVE   CONSTRUCTION.  285 

oil  these  straps  —  that  is  to  say,  we  assume  the  straps  to  be  subjected  to  a  tensile  force 
only,  and  reduce  the  stress  of  10,000  pounds  per  square  inch  of  cross-sectional  area,  so 
as  to  obtain  a  larger  area. 

Hero,  then,  the  question  arises,  How  much  stress  per  square  inch  shall  we  allow? 
For  an  answer  to  this  question  we  must  turn  to  the  straps  at  present  in  use.  In  these 
straps  we  find  that  the  stress  per  square  inch  of  cross-sectional  area  in  the  plane  x  y 
varies,  for  the  rear  straps,  from  6,000  to  7,000  pounds  per  square  inch  ;  and  when  the 
front  straps  are  made  weaker  than  the  rear  ones,  the  stress  per  square  inch  for  the 
former  is  in  the  neighborhood  of  7,000  to  7,500  pounds  per  square  inch.  Our  expe- 
rience leads  us  to  believe  that  a  stress  of  6,500  pounds  per  square  inch  of  cross- 
sectional  area  for  rear  strap,  and  a  stress  of  7,000  pounds  per  square  inch  for  the 
front  one,  both  made  of  the  best  quality  of  hammered  iron,  is  good  practice.  These 
figures  will  be  adopted  in  the  following  calculations.  Hence  we  have  the  following 
rule  : 

KULE  57.  —  To  find  the  thickness  at  g  (Fig.  421)  of  the  rear  strap  on  the  main-rod 
when  the  width  of  the  same  is  given  :  Divide  the  maximum  pressure  on  the  main-rod 
by  6,500  ;  the  quotient  will  be  the  required  number  of  square  inches  in  the  cross- 
sectional  area  at  x  y  (Fig.  422).  From  the  width  of  the  strap  subtract  the  diameter  of 
the  bolt  ;  the  remainder  will  be  the  width  of  the  metal  at  the  weakest  part  of  the 
strap.  Divide  the  required  cross-sectional  area  at  x  y  by  the  width  of  metal  at 
the  weakest  part  of  the  strap  ;  one-half  of  this  quotient  will  be  the  thickness  of  the 
strap  at  //. 

For  finding  the  thickness  of  the  front  main-rod  strap  we  use  the  same  rule,  with 
the  exception  that,  instead  of  dividing  the  maximum  pressure  on  the  main-rod  by 
6,500,  we  divide  it  by  7,000. 

EXAMPLE  92.  —  In  a  locomotive  with  cylinders  18  x  24  inches,  and  a  maximum 
steam  pressure  in  the  cylinders  of  120  pounds  per  square  inch,  we  find  by  the  method 
given  in  Art,  303  that  the  maximum  pressure  on  the  rod  is  31,000  pounds;  the  width 
of  the  front  and  rear  straps  is  2f  inches  ;  the  diameter  of  each  bolt  through  the  straps 
is  if  inch  ;  it  is  required  to  find  the  thickness  of  the  straps  at  g  (Fig.  421). 

The  total  cross-sectional  area  at  x  y  of  the  rear  strap  will  be 

31000 

:=  *-76  square  inches. 


The  width  of  the  metal  at  the  weakest  part  of  the  strap  —  that  is,  at  x  y  —  will  bo 
equal  to 

2?  -  if  =  llf  =  1.8125  inches, 
and 

4.76 

L8125  =   2'62  inches' 
will  be  the  sum  of  the  thickness  of  both  wings  g  g2  of  the  strap.    Therefore, 

2.62 

-^-  =  1.31  inches, 

M 

say  1/g  inches,  is  the  thickness  of  the  rear  strap  at  g. 


286 


MODERN  LOCOMOTIVE   CONSTRUCTION, 


For  the  thickness  of  the  front  strap,  we  have 

31000 

700(7  =  ^'^  S(luare  inches 

in  the  cross-sectional  area  at  x  y. 


and 


2.43 


4.42     _ 
1.8125  ~          ' 

=  1.21  inches,  say  l-^-  inches. 


NOTE. — Some  master-mechanics  make  the  thickness  of  the  front  strap  equal  to 
that  of  the  rear  one ;  therefore,  in  such  cases,  special  calculations  for  the  front  strap 
will  not  be  required. 

311.  Figs.  439,  440  represent  two  views  of  a  main-rod  for  a  consolidation  engine — 
that  is,  an  engine  with  four  pairs  of  drivers,  and  a  pony  truck,  as  shown  in  Fig.  4. 


,_      jw_*r_ 


idf  of  thread 


The  engine  for  which  these  rods  were  designed  has  cylinders  20  inches  diameter  and 
24  inches  stroke,  and  crossheads  like  that  shown  in  Figs.  244,  246,  and  247. 

For  a  high  speed  and  a  maximum  steam  pressure  of  120  pounds  in  the  cylinders, 
the  rear  strap  on  this  rod  is,  in  our  opinion,  too  weak ;  for  the  above  given  maximum 
pressure  and  for  high  speeds,  safer  and  better  results  will  be  obtained  by  making  the 
thickness  of  the  strap  around  the  bolts  at  least  1^  inches ;  and  also  three  }^  inch 
bolts  should  be  used  in  place  of  two. 

SIDE-ROD   STRAPS  FOR  EIGHT-WHEELED  PASSENGER  ENGINES. 

312.  In  comparing  the  side-rod  straps  belonging  to  different  locomotives  of  one 
and  the  same  class  and  size,  but  designed  in  different  locomotive  establishments  or  by 
different  master-mechanics,  we  find  quite  a  variation  in  the  dimensions  of  correspond- 
ing parts  of  the  different  side-rod  straps  around  similar  crank-pins.  This  is,  no  doubt, 
due  to  the  fact  that  it  is  impossible  to  determine  exactly  the  magnitude  of  the  forces 
to  which  the  side-rods  are  subjected. 


MODERX  LOCOMOTirE  CONSTRUCTION.  287 

We  have  already  seen  that  the  main-rods  are  subjected  to  a  tensile  force,  and  also 
to  a  transverse  force  due  to  the  centrifugal  action  at  one  of  its  ends.  The  side- 
rods  are  subjected  to  similar  forces,  but  whose  sum  is  of  less  magnitude  than  that  of 
the  forces  acting  on  the  main-rods.  The  magnitude  of  the  tensile  force  acting  on  the 
main-rod  is  dependent  on  the  whole  pressure  on  the  piston,  whereas  on  side-rods  the 
magnitude  of  the  tensile  force  is  dependent  on  the  weight  on  the  driving  wheels, 
and  partly  also  on  the  condition  of  the  road-bed,  and  on  the  condition  of  the 
working  parts  under  the  engine;  but  the  latter  conditions  will  affect  the  tensile 
force  acting  on  the  side-rods  to  a  greater  degree  than  it  will  affect  the  tensile  forces 
acting  on  the  main-rods.  If,  for  instance,  the  road-bed  is  not  level,  or  the  play 
between  the  axle  box  and  the  wedges  has  been  increased  by  wear,  or  the  engine  is 
running  over  a  sharp  curve,  the  driving  axles,  as  well  as  the  side-rods,  may  at  any 
time  be  thrown  out  of  line  with  each  other,  thereby  suddenly  increasing  the  tensile 
force  on  the  side-rods  on  one  or  the  other  side  of  the  engine ;  and  when,  under  these 
conditions,  the  engine  is  running  at  a  high  rate  of  speed,  this  extra  tensile  force  will 
act  on  the  side-rods  only,  because  these  rods  will  be  affected  by  the  stored-up  energy 
in  the  engine  to  turn  the  wheels,  whereas  the  main-rods  cannot  be  subjected  to  a 
greater  pressure  than  that  due  to  the  pressure  on  the  piston.  Now,  with  these  emer- 
gencies to  provide  for,  it  should  not  be  a  matter  of  surprise  to  hear  of  straps  break- 
ing ;  but  the  frequency  of  such  occurrences  can  be  avoided  by  basing  our  calculation 
on  the  designs  of  such  straps  which  experience  has  taught  us  to  be  correct. 

In  designing  a  side-rod  strap  we  are  guided  by  the  same  reasoning  as  that  given 
in  connection  with  the  design  of  the  main-rod  straps,  and  therefore  conclude  that  the 
weakest  part  of  the  side-rod  strap  is  in  the  plane  x  y  (Fig.  422)  passed  through  the 
axis  of  the  bolt.  Since  the  width  of  the  side-rod  strap  depends  upon  the  dimensions 
of  the  crank-pin,  we  shall  consider  that  the  width  has  been  given,  and  all  that  remains 
for  us  to  determine  is  the  thickness  g  (Fig.  421)  of  the  wing  of  the  strap. 

In  eight-wheeled  passenger  engines  two  driving  wheels  are  connected  on  each 
side  of  the  engine;  consequently,  each  main-rod  in  this  class  of  engines  has  to 
t  ransmit  motion  to  two  driving  wheels ;  and  the  side-rod  has  to  transmit  motion  to 
only  one  driving  wheel.  The  main  driving  wheels  generally  have  to  support  a  greater 
weight  than  the  rear  ones,  as  the  former  have  to  support  a  portion  of  the  weight  of 
the  connecting-rods,  a  portion  of  the  weight  of  the  valve  gear,  and  often  have  a 
heavier  counterbalance  than  the  latter ;  but  for  the  purpose  of  finding  the  proportions 
of  the  side-rod  straps,  we  may  assume  that  each  driving  whool,  whether  front  or  rear, 
supports  an  equal  amount  of  weight.  Again,  the  whole  of  the  total  steam  pressure  on 
the  piston  is  not  utilized  for  the  purpose  of  giving  motion  to  the  driving  wheels ;  yet 
for  the  sake  of  simplicity  we  may  assume  that  such  is  the  case.  Now,  in  order  to 
simplify  matters  still  more,  we  may  again  assume,  as  we  did  in  relation  to  the  main- 
rod,  that  the  side-rods  are  subjected  to  a  tensile  force  only.  Consequently,  since  the 
main-rod  in  eight-wheeled  passenger  engines  has  to  transmit  motion  to  two  driving 
wheels,  and  since  the  side-rod  is,  so  to  speak,  only  a  connecting  link  between  these 
two  drivers,  and  therefore  has  only  to  transmit  motion  to  one  of  them,  it  follows  that, 
under  tin-  foregoing  .assumptions,  the  tensile  force  to  which  the  side-rods  are  subjected 
will  be  equal  to  one-half  the  pressure  on  the  piston.  Comparing  the  cross-sectional 


288  MODERN  LOCOMOTIVE   CONSTRUCTION. 

area  in  the  plane  xy  of  the  side-rod  straps,  at  present  working  successfully,  we  find  that 
the  stress  per  square  inch  of  the  cross-sectional  area  in  the  plane  x  y,  made  by  different 
makers,  vai'ies,  in  round  figures,  from  3,500  to  4,700  pounds.  Our  experience  leads  us 
to  believe  that,  for  eight-wheeled  passenger  engines,  it  is  good  practice  to  make  an 
allowance  for  a  stress  of  4,200  pounds  per  square  inch  of  the  ci'oss-sectional  area 
through  the  weakest  part  of  the  strap  (that  is,  in  the  plane  x  y),  under  the  assump- 
tion that  the  side-rods  are  subjected  to  a  tensile  force  only,  and  that  the  magnitude  of 
this  force  is  equal  to  one-half  the  steam  pressure  on  the  piston ;  this  stress  of  4,200 
pounds  per  square  inch  of  the  weakest  part  of  the  strap  will  be  adopted  in  the  follow- 
ing calculations. 

313.  In  Art.  310  it  will  be  seen  that  for  main-rod  straps  the  stress  allowed  per 
square  inch  in  the  plane  x  y  is  6,500  pounds,  but  for  the  side-rod  straps  we  have  now 
established  a  stress  of  4,200  pounds  per  square  inch  in  a  similar  plane.  This  difference 
is  due  to  the  fact,  as  stated  before,  that  side-rods  are  subjected  to  an  additional  tensile 
force  when  the  engine  is  running  over  uneven  road-beds,  curves,  etc.  Therefore,  in 
order  to  find  the  thickness  at  g  of  the  side-rod  strap  (see  Fig.  421),  we  divide  the  total 
steam  pressure  on  the  piston  by  2 ;  the  quotient  will  be  the  assumed  tensile  force 
acting  on  the  strap ;  dividing  this  quotient  by  4,200,  we  obtain  the  number  of  square 
inches  in  the  smallest  cross-sectional  area  of  the  strap — that  is  to  say,  in  the  plane  x  y. 
Dividing  this  cross-sectional  area  by  the  width  of  the  strap  minus  the  diameter  of  the 
bolt,  we  obtain  the  sum  of  the  thickness  at  g  and  </2 ;  one-half  of  this  sum  will  be  the 
required  thickness  at  g.  This  manner  of  finding  the  required  thickness  can  be  made 
simpler,  hence  the  following : 

RULE  58. — Divide  the  total  steam  pressure  on  the  piston  by  16,800 ;  divide  this 
quotient  by  the  width  of  the  strap  from  which  the  diameter  of  the  bolt  has  been  sub- 
tracted ;  the  last  quotient  will  be  the  thickness  at  g  of  one  of  the  wings  of  the  side- 
rod  strap. 

EXAMPLE  93. — Find  the  thickness  at  g  of  the  side-rod  strap  for  an  eight-wheeled 
passenger  engine ;  cylinders,  17  inches  diameter ;  maximum  steam  pressure  on  the 
piston,  120  pounds  per  square  inch;  width  of  strap,  24  inches;  diameter  of  bolts 
through  the  straps,  £  inch. 

The  total  pressure  on  the  piston  is  equal  to  the  product  of  its  area  in  square  inches 
into  the  steam  pressure  per  square  inch ;  consequently  the  total  pressure  on  the  piston 

will  be 

226.98  x  120  -  27237.60  pounds. 


Dividing  this  pressure  by  16,800,  we  have 

27237.60 


16800 


=  1.621. 


Dividing  this  quotient  by  the  width  of  the  strap  minus  the  diameter  of  the  bolt, 

we  obtain 

1.621 
^~. 1  =  1.17+,  say  li  inches, 

^5    ~    8 

for  thickness  of  the  side-rod  strap  at  g,  Fig.  421. 


MODERN'   LOCOMOTIVE 


289 


:;U.  Iii  Art.  .'!09  wo  have  given  the  reasons  for  the  use  of  two  side-rods 
on  each  side  of  a  Mogul  engine.  For  similar  reasons  we  need  three  side-rods  on 
each  side  of  a  consolidation  engine,  namely,  the  front,  central,  and  rear  side-rods. 


h 


vi-  dla. anttldt  of  tkread. 
t  thread. 


2*q 

U- 

t.-. 

i     T   | 

33  v 

4*r'H  =' 

.  _e  J3£Jj  T 

X  <"<>•  outside  of  thrtad.' 


.  444 


l\dla.  outnlde  of  Ihread^fn 

t  thread*  \_\  ->_\ 


Figs.  441,  442  represent  the  front  side-rod ;  Figs.  44.'!,  444  represent  the  central  side- 
rod  ;  and  Figs.  445,  44ti  represent  the  rear  side-rod  for  a  consolidation  engine  with  cylin- 
ders 20  x  24  inches.  It  will  be  seen  that  the  central  side-rod  connects  two  wheels ;  the 


290 


MODERN  LOCOMOTIVE   COXSTHl'CTIOX. 


straps  on  this  rod  have  forked  ends,  to  which  the  front  and  rear  side-rods  are 
connected.  This  manner  of  connecting  side-rods  differs  from  the  connection  of  the 
side-rods  shown  in  Figs.  437,  438,  in  which  the  rear  side-rod  has  the  forked  end,  and 


r  // 

%  diam.outside  of  thread         ,,  .         ,  *  ~ 

Fig,  445 

42-' . 


Vv— Steel 


the  strap  around  the  main  crank-pin  is  forged  solid.  Again,  in  the  front  and  rear 
side-rods  for  the  consolidation  engine,  provision  has  been  made  to  take  up  the  wear 
of  the  bearings  in  the  ends  of  these  rods  which  are  connected  to  the  central-rod.  This 
makes  a  very  good  arrangement ;  but  it  must  not  be  understood  that  this  design  of 
side-rods  is  always  adopted  in  consolidation  engines,  and  that  the  design  of  side-rods 
shown  in  Figs.  437,  438  is  only  suitable  for  Mogul  engines.  There  are  many  consoli- 
dation engines  in  which  the  side-rods  are  connected  in  a  manner  precisely  similar 
to  that  of  connecting  the  side-rods  in  Mogul  engines. 

315.  In  nearly  all  eight-wheeled  passenger  engines,  4  feet  8.J  inches  gauge,  the 
side-rods  are  placed  outside  of  the  main-rods,  and  therefore  the  brass  bearings,  straps, 
etc.,  at  each  end  of  the  side-rod  will  be  equal  in  size.  In  some  of  the  narrow-gauge 
engines  we  sometimes  find  the  side-rods  plased  inside  of  the  main-rods.  In  such  cases 
the  brass  bearing  at  one  end  of  the  side-rod  will  be  larger  than  that  at  the  other  end 
of  the  same  rod.  Locomotive  builders  sometimes  make  the  straps  for  such  rods  of 
equal  thicknesses.  We  believe  that  in  such  cases  the  better  practice  will  be  to  find 
the  thickness  of  the  strap  at  the  small  end  of  the  rod  by  Rule  58,  and  then  increase 
the  thickness  of  the  strap  at  the  large  end  of  the  side-rod  ten  per  cent.,  provided  the 
widths  of  the  straps  are  equal,  which  is  generally  the  case. 


SIDE-BOD   STRAPS   FOR  MOGUL  ENGINES. 

316.  In  Mogul  engines  the  stress  in  the  side-rods  will — as  in  eight-wheeled  pas- 
senger engines — depend  on  the  weight  on  drivers,  the  condition  of  the  road-bed,  and 
the  condition  of  the  engine,  as  we  have  explained  in  connection  with  side-rods  for 
passenger  engines.  But  in  Mogul  engines  we  have  two  other  elements — not  found  in 
passenger  engines — which  are  detrimental  to  the  strength  of  the  side-rod  strap,  namely, 


UODERX  LOCOMOTITE   CONSTRUCTION. 


291 


Fly. 


•  •  ~Kigi<l-Wheel  Sate 


XOQUZ 


a  larger  number  of  driving  wheels,  and  the  knuckle  joint — that  is,  the  connection  of 
the  two  side-rods  as  shown  at  S  in  Fig.  438. 

317.  Fig.  447  represents  the  wheel  base  of  a  Mogul  engine ;  the  main-  and  side- 
rods  are  indicated  simply  by  their  center  lines.  In  this  class  of  engines  the  main  pair 
of  driving  wheels  B  is  placed  between  the  front  pair  A  and  the  rear  pair  C,  but  is  not 
placed  centrally  between  them.  The  rigid  wheel  base  will  often  depend  on  the  length 
of  the  boiler ;  and  the  distance  between  the  wheels  will  in  many  cases  depend  on  the 
general  design  of  the  boiler 
and  the  valve  gear.  Although 
these  driving  wheels  are  not 
placed  at  equal  distances  apart, 
tin-  arrangement  of  the  equal- 
izing levers  is  such  as  to  throw 
as  nearly  as  possible  the  same 
amount  of  weight  on  each 
driver.  For  the  purpose  of 
designing  the  side-rod  straps,  we  may  assume  that  an  equal  amount  of  weight  is 
placed  on  each  driver.  In  order  to  simplify  the  rule  for  finding  the  thickness  of  the 
strap,  we  shall  again  assume,  as  we  did  in  the  case  of  passenger  engines,  that  the 
whole  steam  pressure  on  the  piston  is  utilized  for  the  purpose  of  giving  motion 
to  the  wheels.  Now,  since  three  drivers  are  connected  on  each  side  of  the  engine, 
and  since  the  same  amount  of  weight  is  assumed  to  be  placed  on  each  driver,  it 
follows  that  one-third  of  the  total  steam  pressure  on  one  piston  will,  according  to 
our  assumption,  be  required  to  turn  one  driving  wheel.  When  the  driving  wheels  are 
turning  in  the  direction  as  indicated  by  the  arrows,  and  the  crank-pins  are  below  the 
centers  of  the  axles,  the  front  side-rod  D  will  be  subjected,  besides  the  transverse 
forces  acting  upon  the  rods,  to  a  tensile  force ;  and  the  rear  side-rod  K  will  be  subjected 
to  a  compressive  force,  each  equal  to  one-third  of  the  pi'essure  on  the  piston,  provided 
the  engine  is  running  over  a  straight  and  perfect  level  road.  On  the  other  hand,  when 
the  crank-pins  are  above  the  center  of  axle,  the  front  side-rod  will  be  subjected,  besides 
the  transverse  forces  acting  upon  the  rods,  to  a  compressive  force  and  the  rear  side-rod 
to  a  tensile  force,  each  again  equal  to  one-third  of  the  pressure  on  the  piston.  The 
exact  magnitude  of  the  transverse  forces  acting  on  the  side-rods  under  the  various 
conditions  of  the  road-bed,  etc.,  cannot  be  determined,  but  the  total  steam  pressure  on 
the  piston  is  known ;  we  therefore  assume,  as  in  the  case  of  passenger  engines,  that 
the  side-rods  are  simply  subjected  to  a  tensile  force  due  to  the  pressure  on  the  piston ; 
design  the  straps  accordingly,  and  make  allowances  for  the  other  forces,  as  will  be 
hereafter  explained.  If  now  we  had  no  transverse  forces  acting  on  the  straps  to 
contend  with,  we  would  simply  divide  one-third  of  the  maximum  piston  pressure  by 
10,000  pounds  (which  is  about  a  fair  allowance  for  stress  per  square  inch  when  we  have 
to  provide  against  a  simple  tensile  force),  and  thereby  obtain  the  number  of  square 
inches  in  the  cross-sectional  area  through  the  weakest  part  of  the  strap.  But  since  we 
have  to  provide  against  the  transverse  forces,  we  must  divide  one-third  of  the 
maximum  steam  pressure  on  the  piston  by  a  number  less  than  10,000,  and  shall  adopt, 
as  before,  4,200.  Hence,  if  we  divide  one-third  of  the  maximum  steam  pressure  on  the 


292  MODERN  LOCOMOTIVE  COXSTRVCTIOX. 

piston  by  4,200,  we  shall  obtain  the  number  of  square  inches  in  the  cross-sectional 
area  through  the  weakest  part  of  the  strap.  Dividing  this  area  by  the  width  of  the 
strap  minus  the  diameter  of  the  bolts  through  it,  and  again  dividing  the  quotient  thus 
obtained  by  2,  we  obtain  a  thickness  for  the  side-rod  straps  which  we  shall  indicate  by 
the  letter  A.  But  this  thickness  A  will  not  be  sufficiently  great,  for  the  following 
reasons :  In  adopting  the  number  4,200,  or,  in  other  words,  allowing  one  square  inch 
for  every  4,200  pounds  of  the  pressure  on  the  strap  due  to  one-third  of  the  maximum 
piston  pressure,  we  provide  for  the  transverse  forces  acting  on  the  strap,  and  also  for 
the  extra  tensile  forces  due  to  the  condition  of  the  road-bed  and  condition  of  the 
engine,  such  as  will  occur  in  passenger  engines.  This  extra  tensile  force,  due  to  the 
condition  of  the  road-bed,  etc.,  will  be  greater  in  Mogul  engines,  on  account  of  the 
increased  number  of  wheels,  than  in  passenger  engines ;  and  besides  this  we  have  the 
knuckle  joint  to  contend  with.  We  therefore  add  to  the  thickness  A  previously  found, 
-nf  of  an  inch  for  the  rear  and  front  strap,  and  add  f  of  an  inch  for  the  central  strap, 
for  all  Mogul  locomotives  having  cylinders  13  inches  diameter  and  upwards.  For 
Mogul  engines  having  cylinders  whose  diameters  are  less  than  13  inches,  add  ^  of  an 
inch  to  the  thickness  A  for  the  front  and  rear  straps,  and  --fa  of  an  inch  to  the  thick- 
ness for  the  central  strap.  These  rules  can  be  stated  in  a  simpler  manner,  as  follows : 

RULE  59. — To  find  the  thickness  g  (Fig.  421)  for  the  side-rod  straps  around  the 
front  and  rear  crank  pins  in  Mogul  engines  having  cylinders  13  inches  diameter  and 
upwards :  Divide  the  total  maximum  steam  pressure  on  one  piston  by  25,200 ;  divide 
the  quotient  thus  obtained  by  the  width  of  the  strap  from  which  the  diameter  of  the 
bolts  through  it  has  been  deducted,  and  add  -f-6-  of  an  inch  to  the  last  quotient ;  the 
sum  will  be  the  required  thickness. 

For  Mogul  engines  having  cylinders  less  than  13  inches  in  diameter,  divide  the 
total  maximum  steam  pressure  on  the  piston  by  25,200;  divide  the  quotient  thus 
obtained  by  the  width  of  the  strap  from  which  the  diameter  of  the  bolts  through  it  has 
been  deducted,  and  add  J  of  an  inch  to  the  last  quotient ;  the  sum  will  be  the  required 
thickness  g  for  the  side-rod  strap  around  the  front  and  rear  crank-pin. 

RULE  GO. — To  find  the  thickness  g  for  the  side-rod  strap  around  the  main  crank- 
pin  in  Mogul  engines  having  cylinders  13  inches  diameter  and  upwards :  Divide  the 
total  maximum  steam  pressure  on  the  piston  by  25,200;  divide  the  quotient  thus 
obtained  by  the  width  of  the  strap  from  which  the  diameter  of  the  bolts  through  it  has 
been  deducted,  and  add  %  of  an  inch  to  the  last  quotient ;  the  sum  will  be  the  required 
thickness. 

For  Mogul  engines  having  cylinders  less  than  13  inches  diameter,  divide  the  total 
maximum  steam  pressure  on  the  piston  by  25,200 ;  divide  the  quotient  thus  obtained 
by  the  width  of  the  strap  from  which  the  diameter  of  the  bolts  through  it  has  been 
deducted,  and  add  -fa  of  an  inch  to  the  last  quotient ;  the  sum  will  be  the  thickness  g 
for  the  central  side-rod  strap. 

EXAMPLE  94. — Find  the  thickness  y  (Fig.  421)  for  the  side-rod  straps  for  a  Mogul 
engine  whose  cylinders  are  18  inches  diameter;  maximum  steam  pressure  on  the 
piston,  120  pounds  per  square  inch ;  width  of  front  and  rear  straps,  2j  inches ;  diame- 
ters of  the  bolts  through  these  straps,  if  inch ;  width  of  central  strap,  2J  inches ; 
diameters  of  bolts  through  the  same,  lr-6-  inches. 


MODERX  LOCOMOTIfK   CONSTRUCTION.  293 

Let  us  first  find  the  thickness  for  the  front  and  rear  straps.  The  total  maximum 
steam  pressure  on  the  piston  is  found  by  multiplying  the  area  of  the  piston  in  square 
inches  by  the  steam  pressure  per  square  inch;  hence  we  have 

254.47  x  120  =  30536.40  pounds. 

30536.40 
25200 

Subtracting  the  diameter  of  a  side-rod  bolt  from  the  width  of  the  strap,  we  obtain 

2.25  -  .9375  =  1.3125  inches, 
and 

1.211 
107  =  .922,  say  |f  of  an  inch, 


which  is  the  thickness  A,  to  which  -n,-  inch  must  be  added.  Hence  jjf  +  ^y  =  l£  inches, 
which  is  the  thickness  g  of  the  front  and  rear  side-rod  straps. 

To  find  the  thickness  g  of  the  central  side-rod  strap,  we  divide,  as  before,  the  total 
maximum  steam  pressure  on  the  piston  by  25,200,  and  obtain  1.211.  This  quotient 
\ve  divide  by  width  of  the  central  strap  minus  the  diameter  of  the  bolt.  We  have 

1.211 

2.5  -  106  =       '  say  ^      an  mc  ' 

which  is  the  thickness  A.  To  this  we  must  add  g  of  an  inch  ;  the  sum  J  +  |  =  li 
inches,  which  is  the  thickness  g  of  the  central  side-rod  strap. 

EXAMPLE  95.  —  What  should  be  the  thickness  of  the  side-rod  straps  for  a  Mogul 
engine  whose  cylinders  are  11  inches  diameter  ;  rear  and  front  straps,  If  inches  wide  ; 
bolts,  £  inch  diameter  ;  central  side-rod  strap,  2  inches  wide  ;  bolts,  f  inch  diameter  ; 
maximum  steam  pressure  on  the  piston,  120  pounds  per  square  inch? 

Total  maximum  steam  pressure  on  the  piston  will  be  equal  to 

95.03  x  120  =  11403.60  pounds, 
and 

11403.60 

252CO 

To  find  the  thickness  g  for  the  front  and  rear  strap,  we  have  first  .452  divided  by 
the  width  of  the  straps  minus  the  diameter  of  the  bolts,  equal  to 

.452 
1.75  -  .75"  ==  •4:>-'  say  *  inch' 

which  is  the  thickness  A.  To  this  we  must,  according  to  rule,  add  \  of  an  inch.  We 
therefore  obtain  i  +  J  =  f  inch  for  the  thickness  g  for  the  front  and  rear  side-rod 
straps. 

For  the  thickness  g  of  the  central  side-rod  straps  we  have 

1  1403.60 

2.VJOO       =  '4>)2' 
and 

.452 
-_>  _   7;,  =  -361,  say  3  inch, 


294 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


which  is  the  thickness  A.     To  this  we  must  add  -fa  inch.     Therefore  t  +  i7^  =  H  inch, 
which  is  the  thickness  y  for  the  central  side-rod  strap. 

318.  For  all  Mogul  engines  in  which  the  maximum  steam  pressure  on  the  piston 
is  120  pounds  per  square  inch,  we  believe  it  is  good  practice  not  to  make  the  thickness 
ff  of  any  side-rod  strap  less  than  2  of  an  inch.  If  our  calculations,  according  to  the 
foregoing  rules,  call  for  a  thickness  less  than  £  inch,  then  the  results  indicate  that 
the  width  of  the  strap  is  excessive,  and  it  should  be  reduced. 


SIDE-ROD   STRAPS   FOR  CONSOLIDATION   ENGINES. 

319.  Fig.  448  represents  the  wheel  base  for  a  consolidation  engine.  Here  we  have 
four  wheels  connected  on  each  side  of  the  engine.  The  wheel  marked  A  is  one  of  the 
first  pair  of  drivers,  B  one  of  the  second  pair,  C  one  of  the  third  pair,  and  I)  one  of 


Fiff.  &4S 


b 

^ Total  TT/teel  Sa.se 

COlf fO  LID  A.TU)Jf  ENGINE 


Fig.  450 

the  rear  or  fourth  pair  of  drivers.  In  some  engines  of  this  class  the  main-rods  are 
connected  to  the  second  pair  of  drjvers.  The  front  side-rod  E,  the  central  side-rod  F, 
and  the  rear  side-rod  £r,  and  also  the  main-rod  M,  are  represented  by  their  center  lines 
only. 


MODERN  LOCOMOTIVE   CONSTRUCTION.  295 

Iii  tliis  class  of  engines  the  equalizing  levers  are  arranged,  as  in  the  former  class, 
to  throw,  as  nearly  as  possible,  an  equal  amount  of  weight  on  each  driver;  hence  for 
the  purpose  of  designing  the  side-rod  straps  we  may  again  assume  that  all  drivers  have 
in  l)i>ar  an  equal  amount  of  weight;  we  may  also  assume,  as  in  the  former  cases,  that 
tlu>  whole  steam  pressure  on  the  piston  is  utilized  for  rotating  the  wheels.  Since  four 
wheels  are  connected  on  each  side  of  the  engine,  it  will  require,  under  the  forego- 
ing assumptions,  one-fourth  of  the  total  steam  pressure  on  the  piston  to  turn  each 
wheel. 

When  the  wheels  are  connected,  as  shown  in  Fig.  448,  the  front  side-rod  E  will 
have  to  transmit  motion  to  one  wheel ;  that  is,  the  wheel  A ;  the  central  side-rod  F 
will  have  to  transmit  motion  to  the  second  wheel  B,  and  by  a  little  reflection  it  will  be 
seen  that  the  same  side-rod  F  has  also  to  transmit  motion,  through  the  front  rod  E,  to 
the  fi-ont  wheel  A.  We  may  therefore  say  that  the  work  performed  by  the  central 
side-rod  /<'  is  equal  to  twice  that  performed  by  the  front  rod  E,  and  conclude  that 
under  our  assumptions  the  tensile  force  acting  on  the  side-rod  F,  due  to  the  steam 
pressure  on  the  piston,  is  equal  to  one-half  of  the  total  maximum  pressure  on  the 
piston ;  and  that  the  tensile  force  acting  on  the  front  side-rod  E  is  eqtial  to  one-fourth 
of  the  total  maximum  steam  pressure  on  the  piston.  Such  conclusion  will  be  correct 
so  long  as  the  engine  is  in  first-class  condition,  and  running  over  a  perfect  and  straight 
road.  But  when  the  engine  is  running  over  curves,  and  the  road-bed  is  not  perfect, 
and  the  wear  has  caused  play  between  the  axle  boxes  and  the  wedges,  the  ratio 
between  the  tensile  forces  acting  on  the  central  and  front  side-rod  will  not  be  exactly 
as  two  to  one. 

Again,  since  the  action  at  one  end  of  the  contra!  side-rod  is  equal  to  the  reaction 
at  the  other  end,  we  conclude  that  the  straps  on  this  rod  should  be  of  equal  dimen- 
sions. Furthermore,  practice  has  shown  that  the  stress  on  the  front  and  rear 
side-rods  is  somewhat  less  than  that  on  the  central  rod.  The  thickness  g  (Fig.  449) 
of  the  side-rod  straps,  obtained  by  the  following  rules,  will  agree  with  modern 
practice. 

RULE  61. — To  find  the  thickness  at  g  (Fig.  449)  for  the  front  and  rear  side-rod 
straps,  for  consolidation  engines :  Divide  the  total  maximum  steam  pressure  on  the 
piston  by  3l3,(iO();  divide  the  quotient  thus  obtained  by  the  width  of  the  strap  minus 
the  diameter  of  the  bolts  through  it ;  add  fa  of  an  inch  to  the  last  quotient ;  the  sum 
will  be  the  required  thickness. 

KULE  62. — For  finding  the  thickness  of  the  central  side-rod  straps,  divide  the 
total  maximum  steam  pressure  on  the  piston  by  33,600;  divide  this  quotient  by 
the  width  of  the  strap,  minus  the  diameter  of  the  bolts  through  it,  and  add  jj  of  an  inch 
to  the  last  quotient ;  the  sum  will  be  the  required  thickness  at  //  (Fig.  449). 

EXAMPLE  96. — What  should  be  the  thickness  of  the  side-rod  straps  for  a  consoli- 
dation  engine  having  cylinders  '20  indies  diameter?  Maximum  steam  pressure  on  the 
piston,  1_()  pounds.  The  width  of  the  central  side-rod  straps,  and  also  the  width  of 
the  front  and  rear  side-rod  straps,  is  2i  inches;  the  diameter  of  the  bolts  through  the 
central  side-rod  is  1|  inches;  the  diameter  of  the  bolts  through  the  front  and  rear 
straps  is  1  inch. 


296  MODERN  LOCOMOTIVE   CONSTBVCT10N. 

The  area  of  a  piston  20  inches  in  diameter  is  314.16  square  inches.  The  total 
maximum  steam  pressure  011  the  piston  will  be 

314.16  x  120  =  37699.20  pounds, 
and 

37699.20 

33600 
The  thickness  at  fj  for  the  front  and  rear  side-rod  straps  will  be  equal  to 

1.122 
tvTirT  +  -fa  =-  lie  inches,  very  nearly. 

The  thickness  for  the  central  side-rod  straps  will  be 

1.122 

2>  _  -j,  +  S  =  1*6  inches,  very  nearly. 

320.  After  the  thickness  at  g  (Fig.  449)  for  the  side-rod  straps  has  been  obtained, 
we  determine  the  thickness  at /and  &  in  a  manner  similar  to  that  adopted  for  finding 
these  thicknesses  for  the  main-rod  straps — namely,  the  thickness  at/  is  made  £  of  an 
inch  greater  than  at  g,  and  the  thickness  at  k  is  made  J  of  an  inch  thicker  than  at  g 
for  small  engines,  and  from  f  to  £  inch  greater  for  large  engines. 

In  some  straps  we  find  the  hole  o  for  the  oil-cup  drilled  through  the  whole  thick- 
ness of  the  wing ;  in  others,  we  find  this  hole  drilled  only  part  way  through  the  wing, 
and  then  a  smaller  hole  s  about  §  of  an  inch  in  diameter  drilled  through  the  remain- 
ing part  of  the  thickness,  as  shown  in  Fig.  449.  The  latter  we  believe  to  be  the 
best  practice,  as  this  will  not  reduce  the  strength  of  the  strap  as  much  as  when  the 
large  hole  is  drilled  clear  through. 


SIDE-ROD   BOLTS. 

321.  In  considering  the  strength  of  the  bolts  which  fasten  the  straps  to  the 
main-rods,  we  have  seen  (Art.  303)  that  the  principal  force  to  which  these  bolts  are 
subjected  is  a  shearing  force.  In  determining  the  number  and  diameter  of  these  bolts 
we  made  their  cross-sectional  area  proportional  to  this  force,  with  certain  allowances  for 
the  other  forces  to  which  they  are  subjected.  It  was  further  seen  (Art.  308)  that  for 
the  light  main-rods  two  bolts  through  each  strap  are  sufficient  for  the  work  they  have 
to  do ;  but  that  for  the  heavy  class  of  main-rods  three  bolts  through  each  strap  must 
be  used,  so  as  to  obtain  the  required  cross-sectional  area,  and  at  the  same  time  avoid 
the  use  of  bolts  of  excessively  large  diameters ;  or,  we  may  say,  bolts  whose  diameters 
are  out  of  proportion  to  other  parts  of  the  rod. 

But  now,  in  considering  the  strength  of  bolts  required  to  hold  the  straps  to 
side-rods,  it  may  be  stated,  at  once,  that  more  than  two  bolts  through  each  side-rod 
strap  are  not  required,  because  two  bolts  will  always  give  us  a  sufficient  cross-sectional 
area  without  using  bolts  of  excessively  large  diameters.  Even  in  the  heaviest  loco- 
motives which  up  to  the  present  time  have  been  built,  it  has  been  found  that  two  bolts 
through  the  side-rod  straps  are  sufficient  to  resist  all  the  forces  to  which  they  are 


LOCOMOTIVE   CONSTRUCTION.  297 

subjected,  and  yet  the  diameters  of  these  bolts  did  not  appear  to  be  too  large,  or,  in 
other  words,  the  bolts  did  not  require  so  much  metal  to  be  drilled  out  of  the  straps  as 
to  increase  the  thickness  of  the  wings  of  the  straps  to  an  unreasonable  extent.  Of 
course,  for  practical  reasons  which  are  obvious,  less  than  two  bolts  through  each  end 
of  the  side-rod  cannot  be  used.  Therefore  we  may  say  that  the  number  of  bolts 
through  each  side-rod  strap  (namely,  two),  for  any  locomotive,  is  established.  In 
order  to  find  the  diameter  of  bolts  for  side-rod  straps,  we  should  know  the  exact  mag- 
nitudes of  the  forces  to  which  they  are  subjected;  but  to  determine  the  magnitude  of 
these  forces  accurately  is  impossible ;  in  fact,  the  same  remarks  made  in  relation  to 
forces  to  which  the  side-rod  straps  are  subjected  are  also  applicable  to  the  forces 
which  tend  to  break  the  bolts  through  the  straps.  We  must  therefore  again  allow 
experience  to  guide  us  in  forming  the  following  rules: 

For  all  practical  purposes  we  may  proceed  in  our  calculations  for  finding  the 
diameters  of  the  side-rod  strap  bolts  as  if  these  bolts  were  subjected  to  a  simple 
shearing  force  only,  due  to  the  weight  on  the  driving  wheels. 

Therefore,  in  the  following  calculations  we  shall  simply  find,  first,  the  diameter  of 
the  bolts  required  to  resist  a  shearing  force,  and  then  add  to  this  diameter  a  certain 
amount  to  allow  for  the  forces  due  to  the  unevenness  of  tracks,  loose  boxes,  etc.,  which 
sum  will  also  be  sufficient  for  the  stress  due  to  the  momentum  of  the  rod. 

322.  Let  us  first  consider  the  diameters  of  bolts  required  for  side-rod  straps  in 
eight-wheeled  passenger  engines.  We  have  already  seen  that,  in  this  class  of  engines, 
there  are  two  driving  wheels  on  each  side,  and  that  the  motion  to  the  rear  driving 
wheel  is  transmitted  through  the  side-rod,  which  connects  the  two.  Now  let  us 
assume,  as  we  have  done  before,  that  the  whole  steam  pressure  on  the  piston  is  utilized 
in  giving  motion  to  the  wheels.  Under  this  assumption,  the  force  acting  in  the 
direction  of  the  length  of  the  side-rod  will  be  equal  to  one-half  the  total  maximum 
steam  pressure  on  the  piston,  and  will  represent  an  assumed  shearing  force,  to  which 
the  bolts  are  subjected. 

Now,  having  established,  for  the  purpose  of  our  calculations,  a  shearing  foi'ce,  we 
next  determine  an  area  proportional  to  this  force,  and  this  may  be  done  in  a  manner 
similar  to  that  adopted  for  finding  cross-sectional  area  of  the  main-rod  strap  bolts, 
namely,  divide  the  pressure  on  the  side-rod,  which  in  this  case  is  assumed  to  be  equal 
to  one-half  of  the  total  maximum  steam  pressure  on  the  piston,  by  8,000;  the  quotient 
will  give  us  an  area  proportional  to  our  assumed  shearing  force;  but  since  this 
area  is  greater  than  that  required  for  the  actual  shearing  force,  we  have  also  made, 
to  a  great  extent,  an  allowance  for  the  force  due  to  the  momentum.  For  the  sake 
of  brevity,  hereafter  we  shall  refer  to  this  area  as  a  simple  shearing  area,  and  the 
force  to  which  it  is  proportioned  the  shearing  force.  Having  found  the  required 
shearing  area,  the  cross-sectional  area  of  one  bolt  is  readily  found.  Since  in  all  side- 
rod  straps  two  bolts  are  used,  and  since  the  shearing  force  tends  to  cut  each  bolt  in 
two  places,  it  follows  thai  the  shearing  area  must  be  equal  to  four  times  the  cross- 
sectional  area  of  one  bolt;  hence  we  divide  the  former  area  by  4;  the  quotient  will  be 
the  cross-sectional  area  in  inches  of  one  bolt ;  the  diameter  corresponding  to  the  last 
area  will  be  the  diameter  of  the  bolt  required  to  resist  the  shearing  force.  In  order  to 
allow  for  the  forces  due  to  uneven  tracks,  etc.,  we  must  add  £  of  an  inch  to  the  diame- 


298  MODERN  LOCOMOTIVE   CONSTRUCTION. 

ter  thus  found ;  the  sum  will  be  the  required  diameter  of  the  side-rod  bolts  in  eight- 
wheeled  passenger  engines.  This  rule  can  be  stated  in  a  simpler  manner,  as  we  shall 
presently  show. 

In  Mogul  and  ten-wheeled  engines  we  have  three  driving  wheels  connected  on  each 
side  of  the  engines ;  we  therefore  divide  one-third  of  the  maximum  steam  pressure  on 
the  piston  by  8,000 ;  the  quotient  will  be  the  shearing  area  in  square  inches ;  dividing 
this  area  by  4,  we  again  obtain  the  cross-sectional  area  of  one  bolt,  the  diameter  of 
which  must  also  be  increased  by  some  given  amount  to  resist  the  additional  forces 
which  come  into  play.  This  additional  amount  will  be  given  in  the  rules  which  are 
to  follow. 

The  diameters  of  the  bolts  through  side-rod  straps  in  consolidation  engines  are 
found  in  a  manner  similar  to  the  foregoing.  In  these  engines  we  have  four  wheels, 
connected  on  each  side,  consequently  we  divide  one-fourth  of  the  maximum  steam 
pressure  on  the  piston  by  8,000,  so  as  to  obtain  the  shearing  area  of  the  bolts  through 
the  front  and  rear  straps ;  dividing  the  latter  by  4,  we  obtain  the  cross-sectional  area 
of  one  bolt,  whose  corresponding  diameter  must  again  be  increased  by  a  certain 
amount,  as  will  be  given  in  the  following  rules : 

RULE  63. — To  find  the  diameter  of  a  side-rod  strap  bolt  for  an  eight- wheeled  pas- 
senger engine :  Divide  the  total  maximum  steam  pressure  on  the  piston  by  G4,000 ;  the 
quotient,  will  be  the  cross-sectional  area  of  one  bolt  required  to  resist  the  sheai'ing  force. 

To  the  corresponding  diameter  of  this  area  add  ^  of  an  inch ;  the  sum  will  be  the 
required  diameter  of  the  bolts. 

EXAMPLE  97. — The  diameter  of  the  cylinders  of  an  eight-wheeled  passenger  engine 
(such  as  is  shown  in  Fig.  1)  is  18  inches ;  maximum  steam  pressure  per  square  inch  of 
piston,  120  pounds ;  find  the  diameter  of  the  side-rod  bolts. 

The  total  maximum  steam  pressure  on  the  piston  is  equal  to  its  area  multiplied 
by  120.  The  area  of  an  18-inch  piston  is  equal  to  254.47  square  inches ;  and  254.47  x 
120  =  30536.4  pounds. 

30536.4 
64000     =0.4/7  square  inch. 

The  nearest  diameter  of  a  bolt  corresponding  to  a  cross-sectional  area  of  0.477 
square  inch  is  ||  inch;  neglecting  the  ^  of  an  inch  and  adding  £  of  an  inch 
to  the  diameter  found,  we  have  f  +  £  =  J  inch,  which  is  the  required  diameter  of  the 
side-rod  strap  bolts  for  eight-wheeled  passenger  engines. 

RULE  64. — To  find  the  diameters  for  the  side-rod  strap  bolts  in  a  Mogul  engine : 
Divide  the  total  maximum  steam  pressure  on  the  piston  by  96,000 ;  the  quotient  will 
be  the  cross-sectional  area  in  square  inches  of  one  bolt  necessary  to  resist  the  shearing 
force.  To  the  corresponding  diameter,  which  we  designate  by  the  letter  A,  add  £  of  an 
inch ;  the  sum  will  be  the  required  diameter  of  the  bolts  through  the  straps  around  the 
front  and  i%ear  crank-pin.  Again,  to  the  diameter  A  add  f  of  an  inch ;  the  sum  will  be 
the  required  diameter  of  the  bolts  through  the  straps  around  the  central  or  main 
crank-pin. 

EXAMPLE  98. — Find  the  diameters  for  the  side-rod  bolts  in  a  Mogul  engine,  having 
cylinders  19  inches  in  diameter ;  maximum  steam  pressure  per  square  inch  of  piston, 


MODEBX  LOCOMOTIVE   CO\STJIUCT1(>.\.  299 

120    pounds.     The    area   of  a    19-inch   piston   is   283.53   square    inches;    the   total 
maximum  pressure  on  the  piston  will  be  equal  to  283.53  x  120  =  34,023.6  pounds. 

34023.6 
,,  .  =  0.3o4  square  inch. 

The  nearest  diameter  corresponding  to  an  area  of  0.354  square  inch  is  •$-£ 
inch.  Therefore  1«  +  i  =  H  inch,  which  is  the  diameter  for  the  bolts  through  side- 
rod  straps  for  the  front  and  rear  crank-pin.  And  }$  +  I  =  IMJ  inch,  which  is  the 
diameter  of  the  bolts  through  the  side-rod  strap  for  the  central  or  main  crank-pin. 

RULE  65. — To  find  the  diameters  for  the  side-rod  strap  bolts  in  a  consolidation 
engine:  Divide  the  total  maximum  steam  pressure  on  the  piston  by  128,000;  the 
ijuotient  will  be  the  cross-sectional  area  in  square  inches  of  each  bolt  through  the 
front  and  rear  straps  necessary  to  resist  the  shearing  force.  To  the  corresponding 
diameter,  which  we  shall  designate  by  the  letter  B,  add  §  of  an  inch ;  the  sum  will  be 
the  required  diameter  of  the  bolts  through  the  straps  for  the  front  and  rear  crank- 
pins.  Again,  to  the  diameter  B,  add  £  inch;  the  sum  will  be  the  diameter  of  the 
bolts  through  the  straps  for  the  second  and  third  crank-pins. 

EXAMPLE  99. — Find  the  diameters  for  the  side-rod  bolts  in  a  consolidation  engine 
having  cylinders  20  inches  in  diameter ;  maximum  steam  pressure  per  square  inch  of 
piston,  120  pounds. 

The  area  of  a  20-inch  piston  is  equal  to  314.16  square  inches ;  the  total  maximum 
steam  pressure  on  the  piston  will  be  equal  to  314.16  x  120  =  37699.2  pounds. 

37699.2 

1^8000  "  square  inch. 

The  nearest  diameter  corresponding  to  an  area  of  0.294  square  inch  is  § 
inch.  Therefore  f  +  f  =  1  inch,  which  is  the  diameter  of  the  bolts  through  the 
straps  around  the  front  and  rear  crank-pins.  And  £  +  £  =  l£  inches,  which  is  the 
diameter  of  the  bolts  through  the  straps  for  the  second  and  third  crank-pins. 

323.  The  bolts  through  the  side-rod  and  also  through  the  main-rod  straps  should 
be  placed  as  close  to  the  keys  as  possible,  leaving  only  sufficient  room  to  tighten  the 
nuts.    The  distance  rf,  Fig.  449  or  Fig.  451,  that  is,  the  distance  from  end  of  the  wing 
of  the  strap  to  the  first  bolt,  is  generally  made  equal  to  about  one  and  a  half  times  the 
diameter  of  the  bolt.     The  distance  from  center  to  center  of  bolts  varies  from  2  to  3 
inches,  depending  on  the  diameters  of  the  bolts ;  for  engines  with  cylinders  16  inches 
in  diameter  and  upwards,  this  distance  is  generally  3  inches,  and  for  engines  having 
cylinders   10  or  11   inches,  it    is   •_'   indies,  and  in  some  cases  even  less  than  that. 
Good  practice  seems  to  indicate  that  these  bolts  should  be  as  close  to  each  other  as  a 
sufficient  clearance  for  the  wrench  will  allow. 

The  side-rod  bolts  are  generally  tapered,  the  taper  varying  from  ^  to  £  of  an  inch 
in  12  inches.  The  diameters  of  the  bolts  found  in  the  foregoing  calculations  are  the 
small  diameters. 

324.  The  cross-section  of  the  keys  is  usually  rectangular.     We  believe  the  better 
practice  will  be  to  round  off  the  side  H,  as   shown   in   Figs.   449,  450,  because   it 
has  been  found  that,  when  keys  of  rectangular  cross-section  are  used,  the  strap  is 


300 


MODERN  LOCOMOTIVE    CONSTRUCTION. 


liable  to  crack,  as  indicated  at  p  in  Fig.  450,  which  shows  that  sharp  corners  at  these 
points  impair  the  strength  of  the  strap. 

The  thickness  of  the  key  is  generally  made  from  £  to  \  of  an  inch  less  than  the 
diameters  of  the  bolts.  The  width  a  at  the  small  end  of  the  key  (Figs.  449,  451)  varies 
from  one  to  one  and  a  half  times  the  diameter  of  the  bolts ;  the  latter  is  preferable 
for  main-rods,  the  former  for  side-rods.  The  length  /  of  the  key  is  usually  made 
equal  to  one  and  a  half  times  the  total  width  of  the  strap. 

The  taper  of  these  keys  varies  from  f  to  Ij  inches  in  12  inches ;  the  former  is  the 
most  common. 

Although  sometimes  the  keys  are  made  of  wrought-iron,  the  best  practice  is  to 
make  them  of  steel. 

325.  Figs.  451,  452  represent  a  main-rod  whose  front  end  is  designed  to  connect 
to  a  crosshead  working  between  two  slides,  similar  to  the  one  shown  in  Figs.  241,  243. 


1 — -«*'- — i tat 


-•* —  -v-s 


The  key  in  the  front  end  is  inserted  horizontally,  because  there  is  not  sufficient  room 
between  the  projecting  ends  of  the  crosshead  to  place  the  key  vertically.  The  key  is 
necessarily  made  short,  so  as  to  clear  other  parts  of  the  machinery.  It  bears  against  a 
cast-iron  block  (7,  which,  when  the  key  is  drawn  in,  forces  the  brasses  B  B2  against 
each  other.  The  bolt  D  simply  prevents  the  cast-iron  block  from  slipping  out  of 
position. 


MODERN  LOCOUOTirE   CONSTRUCTION. 


301 


FORMS   OF  BODS. 

326.  The  favorite  locomotive  main-rod  is  the  solid  one  of  rectangular  form — that 
is  to  say,  a  rod  whoso  tranverse  section  is  similar  to  that  shown  in  Fig.  453 ;  they  are 
made  stiff  enough  so  as  not  to  buckle. 

Main-rods  are  subjected  alternately  to  a  tensile  and  compressive  force,  due  to  the 
steam  pressure  on  the  piston;  the  intensity  of  these  forces  is  increased  by  the 
obliquity  of  the  rods,  and  also  by  positions  of  some  of  the  mechanism  which  at  times 
must  work  out  of  alignment  when  the  engine  is  running  over  curves  or  uneven  tracks. 
The  compressive  force  has  a  tendency  to  buckle  the  rods  in  the  direction  of  arrow  2 
(Fig.  457) ;  in  order  to  prevent  buck- 
ling in  this  direction,  a  definite  thick- 
ness at  C  and  D  (Fig.  457)  will  bo 
required ;  in  all  main-rods  this  thick- 
ness is  uniform  throughout.  The 
main-rods  are  also  subjected  to  a 
transverse  force,  due  to  the  centrip- 
etal acceleration,  and  this  force  has 
a  tendency  to  bend  or  break  the  rod 
in  the  direction  of  arrow  3  (Fig.  456) ; 
to  prevent  any  change  of  form  in  this  direction,  a  definite  depth  at  A  and  B  will 
be  required ;  this  depth  increases  uniformly  from  A  to  B. 

The  thickness  and  depth  of  main-rods,  made  by  different  builders  and  master- 
mechanics,  for  the  same  class  and  size  of  locomotives,  vary  somewhat,  but  the  average 
of  good  practice  seems  to  point  to  the  following  proportions : 

In  the  smallest  transverse  section  of  a  main-rod,  that  is,  at  A,  the  depth  /  h, 
Fig.  453,  should  be  equal  to  If  times  the  thickness  fg ;  the  depth  at  B,  Fig.  456, 
should  be  25  per  cent,  greater  than  that  at  A  ;  and  furthermore,  the  area  of  the  trans- 
verse section  at  A  should  be  such  as  to  contain  1  square  inch  for  every  5,000  pounds 


Fig.  453 


fill.  454 


Fig.  455 


I 

Fig.  450 


Fig.  457 

of  the  total  maximum  steam  pressure  on  the  piston.  Hence,  within  the  limits  of 
ordinary  locomotive  practice,  we  may  adopt  the  following  rules,  which  arc  based  upon 
the  proportions  just  given. 

RULE  66. — To  find  the  area  of  the  smallest  transverse  section  of  a  main-rod,  that 
is,  the  area  of  the  cross-section  at  A,  Fig.  45(i,  divide  the  total  maximum  steam  press- 


302  MODERN  LOCOMOTIVE   CONSTRUCTION. 

ure  on  the  piston  by  5,000 ;  the  quotient  will  be  the  number  of  square  inches  in  the 
area  of  the  transverse  section  at  A. 

EXAMPLE  100. — The  maximum  steam  pressure  on  the  piston  of  a  consolidation 
engine  is  120  pounds  per  square  inch,  cylinders  20  inches  diameter ;  what  should  be 
the  number  of  square  inches  in  transverse  section  through  the  smallest  part  of  the 
main-rod! 

The  area  of  a  20-inch  piston  is  314.16  inches ;  the  maximum  steam  pressure  on 
the  piston  will  be  equal  to 

314.16  x  120  =  37699.20  pounds, 

and  the  number  of  square  inches  required  in  the  area  of  the  smallest  cross-section  of 
the  main-rod  will  be  equal  to 

37699.20 
rnnn       '  ««Si  square  inches. 

RULE  67. — To  find  the  thickness  and  depth  of  the  main-rod  at  its  smallest  transverse 
section,  that  is,  at  A,  Fig.  456,  and  also  the  depth  of  the  rod  at  B,  multiply  the  cross- 
sectional  area  of  the  smallest  section  of  the  rod,  as  found  by  Eule  66,  by  4,  and  divide 
the  product  by  7 ;  the  square  root  of  the  quotient  will  be  the  required  thickness  C  or 
D  (Fig.  457)  in  inches. 

To  find  the  depth  of  the  rod  at  A,  multiply  the  thickness  of  the  rod  in  inches  by 
1.75 ;  the  product  will  be  the  depth  in  inches  at  A.  For  the  depth  at  B,  increase  the 
depth  of  A  by  25  per  cent. 

EXAMPLE  101. — What  should  be  the  thickness  and  depth  of  a  main-rod  for  a 
consolidation  engine  having  cylinders  20  inches  diameter,  steam  pressure  on  piston 
120  pounds  per  square  inch  ? 

Here  we  must  first  find  the  area  in  square  inches  of  the  smallest  cross-section  of 
the  rod,  that  is,  at  A.  In  the  last  example  we  found  this  area  to  be  7.53  inches. 
Hence,  the  thickness  of  the  rod  will  be  equal  to 


7.53  X  4        0  nv  •     i 

-  =  2.07  inches. 

7 

The  depth  of  the  rod  at  A  will  be  equal  to  2.07  x  1.75  =  3.62  inches. 

The  depth  at  B  will  be  equal  to  3.62  x  1.25  =  4.52  inches. 

Avoiding  fractions  less  than  ^  inch,  we  find  that  the  thickness  of  this  rod  should 
be  2116  inches;  the  depth  at  A,  3jj  inches;  and  the  depth  at  B,  4^  inches. 

EXAMPLE  102. — What  should  be  the  thickness  and  depth  of  a  main-rod  for  an 
eight-wheeled  passenger  engine  having  cylinders  18  inches  diameter,  maximum  steam 
pressure  in  cylinder  140  pounds  per  square  inch  ? 

The  total  maximum  steam  pressure  on  the  piston  is  equal  to  254.47  x  140  = 
35625.8  pounds. 

According  to  Eule  66,  the  area  of  the  smallest  cross-section  should  be  equal  to 

35625.8 

~~k7vvr~    =  '-12+  square  inches ; 

and  according  to  Eule  67,  the  thickness  of  the  main-rod  will  be  equal  to 


7.12x4 


XODERX  LOCOMOTIVE  CONSTRUCTION.  303 

The  depth  of  the  rod  at  A  will  be  equal  to 

2.02  x  1.75  =  3.53  inches. 

The  depth  of  the  rod  at  B  will  be  equal  to  3.53  x  1.25  =  4.41  inches. 
Avoiding  fractious  less  than  ^  inch,  the  thickness  of  the  rod  will  be  2  inches ;  depth 
at  A,  3J  inches;  and  depth  at  B,  4,^  inches. 

It  will  be  noticed  in  these  examples  that  in  the  small  tranverse  section  of  the  rod 
tin1  side /A,  Fig.  453,  is  If  longer  than  the  side  f  <j.  Should  it  be  required  to  have  a 
different  ratio  between  these  sides,  say  that  /  h  shall  be  If  longer  than  /  <?,  then 
multiply  the  area  in  square  inches,  as  found  by  Rule  66,  by  8,  and  divide  the  product 
by  13 ;  the  square  root  of  the  quotient  will  be  the  thickness ;  and  the  thickness  thus 
found  multiplied  by  If,  or  1.625,  will  give  the  depth  of  the  rod  at  A. 

The  foregoing  rules  are  applicable  to  locomotive  main-rods  only;  and  even  in 
locomotive  practice,  these  rules  will  give  satisfactory  results  only  so  long  as  the  length 
of  the  main-rod  is  not  greater  or  much  greater  than  60  times  the  width  of  the  rod 
found  by  these  rules.  Main-rods  whose  lengths  are  greater  than  60  times  the  thick- 
ness, or  connecting-rods  for  other  engines  in  which  this  ratio  and  the  maximum  steam 
pressure  on  the  piston  differs  greatly  from  ordinary  locomotive  practice,  should  be 
treated  as  upright  columns  or  pillars,  with  rounded  or  pointed  ends  supporting  a  load ; 
and  the  dimensions  of  these  rods  should  be  found  by  rules  given  in  books  treating  on 
the  strength  of  materials. 

Sometimes  we  find  locomotive  main-rods  with  the  edges  chamfered,  whose  cross- 
sections  will  appear  as  shown  in  Fig.  454.  This  we  believe  to  be  bad  practice,  as  it 
ruts  the  metal  away  in  places  where  it  is  needed  the  most.  Chamfered  edges  do  not 
add  to  the  beauty  of  a  rod,  but  unnecessarily  increase  the  expense  of  making  them, 
and  when  done  no  advantage  whatever  is  gained.  The  best  practice  is  simply  to  take 
off  the  sharp  corners  to  as  small  degree  as  possible,  so  as  to  prevent  a  person  cutting 
liis  hands  or  otherwise  hurting  himself  in  oiling,  cleaning,  or  inspecting  a  locomotive. 

Sections  of  rods,  as  shown  in  Fig.  455,  should  also  be  avoided,  because  such  forms 
only  impair  the  strength  of  the  rod  without  gaining  any  advantage. 

327.  Main-rods  are  often  made  of  iron,  sometimes  of  steel ;  when  of  the  former, 
the  bt-st  quality  of  hammered  iron  must  be  used.  The  rules  for  finding  the  dimensions 
of  main-rods  (given  in  Art.  326)  are  suitable  for  rods  made  of  the  best  hammered  iron. 
When  made  of  steel  they  may  be  made  slightly  lighter.  But  since,  in  many 
cases,  steel  rods  are  adopted  simply  for  the  purpose  of  ensuring  greater  safety,  and  not 
so  much  for  the  purpose  of  reducing  the  weight,  no  difference  in  the  dimensions 
between  iron  and  steel  rods  is  made. 

For  the  sake  of  convenience  in  designing,  we  have  given  the  following  tables 
containing  the  dimensions  for  iron  main-rods.  These  have  been  determined  by  the 
rules  given  in  Art.  326.  In  the  dimensions  given  we  have  avoided  those  containing 
less  than  one  -j^  of  an  inch,  and  selected  such  as  agreed  nearest  with  the  decimals 
found. 


304 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


TABLE    22. 

THICKNESS    AND  DEPTH    OF    MAIN-RODS    AT    THE    SMALL    AND    LARGE    ENDS.      RODS  MADE  OF    BEST 
HAMMERED   IRON.      MAXIMUM   STEAM   PRESSURE   PER  SQUARE  INCH   OF   PISTON,  120  POUNDS. 


Diameter  of  Cylinders. 

Thickness. 

Depth  at  Small  End. 

Depth  at  Large  End. 

9" 

H" 

If" 

2" 

10" 

i" 

1-ft" 

2*" 

11" 

•  H" 

2" 

24" 

12" 
13" 

H" 
if" 

2A" 

2|" 

2f" 

aft" 

14" 
15" 
16" 

iV 

ift" 
14" 

2ft" 
2J" 
2J" 

3,%" 
3|" 

3f" 

17" 

if" 

»V 

31" 

18" 

u« 

3J" 

W 

19" 

2" 

W' 

4±" 

20" 

2ft" 

3|" 

44" 

21" 

2,V 

3||" 

4f" 

22" 

2ft" 

4" 

5" 

TABLE   23. 

THICKNESS    AND    DEPTH    OF    MAIN-RODS    AT    THE    SMALL    AND    LARGE    ENDS.      RODS   MADE   OF   BEST 
HAMMERED  IRON.      MAXIMUM  STEAM   PRESSURE   PER  SQUARE   INCH  OF   PISTON,   130   POUNDS. 


Diameter  of  Cylinders. 

Thickness. 

Depth  at  Small  End. 

Depth  at  Large  End. 

9" 

1" 

Itt" 

2i" 

10" 

U" 

•    ir 

2|" 

11" 

1ft" 

aA" 

2ft" 

12" 

1ft" 

2i" 

2[f" 

13" 

1ft" 

2|" 

3,V 

14" 

14" 

2J" 

3±" 

15" 

If" 

2jg" 

3i" 

16" 

1J" 

3" 

3i" 

17" 

iff" 

3ft" 

4" 

18" 

iH" 

3f" 

4i" 

19" 

*S" 

3ft" 

4i" 

20" 

24" 

3f" 

4H" 

21" 

2±" 

3}g" 

4^|" 

22" 

2f" 

4*" 

5ft" 

TABLE  24. 

THICKNESS    AND    DEPTH  OF    MAIN-RODS  AT    THE    SMALL    AND    LARGE    ENDS.      RODS   MADE   OF    BEST 
HAMMERED   IRON.     MAXIMUM  STEAM  PRESSURE  PER  SQUARE   INCH  OF  PISTON,   140   POUNDS. 


Diameter  of  Cylinders. 

Thickness. 

Depth  at  Small  End. 

Depth  at  Large  End. 

9" 

1" 

It" 

2ft" 

10" 

H" 

iff" 

2ft" 

11" 

i±" 

2*" 

2H" 

12" 

ift" 

2f" 

2W 

13" 

i,y 

2ft" 

3ft" 

14" 

i  ;••„" 

2f" 

3ft" 

15" 

1M" 

211" 

3-fI" 

16" 

I«" 

3i" 

3H" 

17" 

lit" 

3ft" 

4i" 

18" 

2" 

3i" 

4|" 

19" 

2i" 

3H" 

4«" 

20" 

24-" 

3^" 

4|" 

21" 

2f" 

4i" 

54" 

22" 

21%" 

4ft" 

5|" 

\ioi>j-:i;.\- 


COSSTRVCTION. 


305 


TABLE  25. 

THICKNESS    AND    DEPTH    OF  MAIN-BODS    AT    THE   SMALL    AND    LARGE    ENDS.      RODS   MADE   OF    BEST 
HAMMERED  IRON.      MAXIMUM   STEAM  PRESSURE   PER  SQUARE   INCH  OF  PISTON,  150  POUNDS 


Diameter  of  Cylinders. 

Thickness. 

Depth  at  Small  End. 

Depth  at  Large  End. 

9" 

iiV' 

US" 

2i" 

10" 

H" 

2" 

2J" 

11" 

U" 

»" 

2f" 

12" 

If" 

•2A" 

3" 

13" 

H" 

2»" 

3J" 

14" 

14" 

art" 

31" 

15" 

It" 

3" 

3f" 

16" 

if" 

3i" 

4" 

17" 

W" 

3,V 

4±" 

18" 

-',',," 

31" 

*ft" 

19" 

2ft" 

8J" 

4i  1" 

20" 

2ft" 

4ft" 

5ft" 

21" 

2,76" 

4i" 

5  A,-" 

22" 

-',",," 

4ft" 

5ft" 

TABLE  26. 

THICKNESS   AND    DEPTH    OF    MAIN-RODS  AT    THE    SMALL    AND    LARGE  ENDS.      RODS  MADE    OF    BEST 
HAMMERED  IRON.      MAXIMUM   STEAM   PRESSURE   PER  SQUARE   INCH  OF  PISTON,  100  POUNDS. 


Diameter  of  Cylinders. 

Thickness. 

Depth  at  Small  End. 

Depth  at  Large  End. 

9" 

It's" 

11" 

2ft" 

10" 

1ft" 

2ft" 

2f" 

11" 

1ft" 

2ft" 

2?" 

12" 
13" 

1ft" 

i,y 

84" 

2ti" 

3i" 
3f" 

14" 

Hi" 

2ir 

3J" 

15" 

Hi" 

3i" 

air 

16" 

MS" 

3ft" 

4ft" 

17" 

2ft" 

3ft" 

4,'," 

18" 

2ft" 

3f" 

4J" 

19" 

2ft" 

3  It" 

*«" 

20" 

2f" 

4ft" 

5i" 

21" 

21" 

4|" 

5i" 

22" 

2i" 

4$" 

51" 

SIDE-KODS. 

328.  When  we  consider  all  the  conditions  under  which  a  side-rod  must  transmit 
motion  from  one  wheel  to  another,  it  will  be  seen  that  to  design  such  a  rod  is  not  as 
easy  as  to  design  a  main-rod.  We  have  seen  that  a  main-rod  should  be  stiff  enough  to 
do  its  work  without  buckling  in  any  direction ;  and  since  we  can  estimate  very  closely 
the  pressure  to  which  it  will  be  subjected,  its  strength  to  resist  these  pressures  can  be 
readily  determined,  as  shown  by  the  foregoing  rules.  Side-rods,  however,  labor  under 
disadvantages  to  which  the  main-rods  are  not  subjected.  Wear  will  create  play 
between  the  axle  boxes  and  wedges,  allowing  the  axles  to  run  out  of  their  proper 
adjustment,  thereby  throwing  an  extra  stress  on  the  side-rods ;  uneven  tracks  will 
throw  the  side-rods  out  of  parallelism,  which  will  again  increase  the  stress  on  the  rods ; 
unequal  wear  of  tires,  which  practically  means  a  difference  in  the  diameter  of  the 


306  MODERN  LOCOMOTIVE   CONSTRUCTION. 

wheels,  and  consequently  that  one  or  the  other  wheel  must  slip  a  certain  amount 
during  each  revolution;  but  this  slip,  due  to  the  unequal  diameters  of  the  wheels, 
cannot  take  place  through  any  other  agency  than  the  side-rod,  and  consequently  the 
rod  will  again  be  subjected  to  an  extra  thrust  or  pull.  But  slip  is  not  only  due  to  the 
unequal  wear  of  tires ;  it  is  also  caused  by  the  form  of  the  tread  of  tires,  and,  as  we 
have  seen  in  previous  articles,  many  tires  have  a  cone  tread ;  consequently,  in  curving, 
the  wheels  having  such  treads  will  run  on  one  side  of  the  engine  on  larger  diameters 
than  on  the  other  side,  and  consequently  slip  must  occur.  Another  feature  which 
presents  itself  in  connection  with  curving  is  that  the  play  between  the  axle  boxes  and 
wedges  will  cause  the  axles  to  run  out  of  parallelism,  and  all  this  tends  to  throw  extra 
stress  on  the  side-rods.  Comparatively  sudden  stopping  by  the  application  of  brakes, 
running  over  slippery  places  on  the  rails,  or  incautious  use  of  sand  often  plays 
mischief  with  the  side-rods. 

Now,  these  conditions  are  the  disadvantages  under  which  a  side-rod  labors,  and 
may  at  times  throw  on  it  extraordinary  pressures  which  cannot  be  accurately  deter- 
mined, but  can  only  be  appi'oximately  estimated. 

Side-rods  should  be  made  as  light  as  possible,  so  as  to  reduce  the  stresses  due  to 
the  action  of  the  centrifugal  force  to  a  minimum,  yet  they  must  be  strong  enough  to 
resist  the  tensile  and  compressive  forces  to  which  they  are  alternately  subjected. 
When  side-rods  are  subjected  to  a  compressive  force  or  thrust,  they  must  not  buckle 
in  a  vertical  direction,  that  is,  in  the  direction  of  arrow  2,  Fig.  458 ;  yet,  under  certain 
circumstances,  it  is  desirable  that  they  should,  to  a  limited  extent,  slightly  spring  or 
buckle  in  a  horizontal  direction,  that  is,  in  the  direction  of  arrow  3.  The  reason  for 
desiring  a  slight  spring  of  the  side-rods  in  a  horizontal  direction  is  to  obtain  a  certain 


I 

Fig.  458 

i 

Ifi 

*' 

Fig.  459 

amount  of  flexibility,  so  as  to  avoid  excessive  jerks  on  the  rod  and  crank-pin,  and 
thereby  lessen  the  danger  of  heating  the  crank-pin  brasses,  or  breaking  the  crank-pins 
or  side-rods.  Here  we  notice  a  difference  between  the  essential  conditions  demanded 
from  a  main-rod  and  a  side-rod ;  the  former  must  do  its  work  without  buckling  in  any 
direction,  the  latter  should  not  buckle  in  a  vertical  direction,  but  should  have  a 
certain  amount  of  flexibility  by  springing  to  a  slight  extent  in  a  horizontal  direction ; 
and  these  requirements  the  designer  should  not  lose  sight  of. 

Now,  a  side-rod  which  shall  meet  all  the  demands  made  upon  it  must  have  a 
proper  distribution  of  metal,  and  must  also  be  of  such  a  form  as  will  reduce  the  cost  of 
manufacture  to  a  minimum.  Consequently,  various  types  of  side-rods  are  now  in  use. 

In  the  early  stages  of  locomotive  building,  many  side-rods  with  circular  transverse 
sections  were  used,  the  diameter  of  the  central  section  being  larger  than  the  diameters 


M(>l>i:i;\    LOCOMOTIl'K    fOXSTRVCTION. 


307 


of  the  end  sections.  This  type  of  rod  was  finally  abandoned,  because,  although  cheap 
to  manufacture,  it  had  the  same  rigidity  vertically  and  laterally,  which,  as  we  have 
seen,  is  objectionable. 

Probably  the  most  popular  type  of  side-rod  at  present  in  use  is  that  shown  in 
Figs.  458,  459.  This  rod  has  a  uniform,  rectangular  tranverse  section  throughout.  It 
1ms  a  certain  amount  of  flexibility  in  the  direction  of  arrow  3,  and  is,  comparatively, 
not  a  costly  rod  to  make.  This  rod  is  extensively  used  both  for  freight  and  passenger 
engines.  On  account  of  its  cheapness,  it  is  nearly  always  adopted  for  freight  engines ; 


I 

Fig.  46O                                                                                          \\ 

1 

|  <r 

' 

; 

fig.  461 


Fig.  462 


Fig.  463 

but  for  fast  passenger  service  its  metal  is  not  considered  to  be  correctly  distributed, 
and,  consequently,  we  now  frequently  meet  with  passenger  engines  for  which  a  type 
of  side-rod  such  as  is  shown  in  Figs.  460,  461  has  been  adopted.    The  transverse 
section  of  this  rod  is  of  the  I  form,  as  shown  in  Fig.  464,  drawn  to 
a  larger  scale  than  Figs.  460,  461.     For  some  rods  the  section  is 
made  uniform  throughout ;  in  others,  it  is  deeper  at  the  center  than 
at  the  ends.     The  advantage  of  this  form  of  rod  is  that,  with  an 
amount  of  metal  equal  to  that  used  for  a  rod  with  a  solid  rectangular 
section,  its  depth   can  be  made  greater  than  the  depth  of  the  latter, 
and  consequently  it  is  stronger  to  resist  the  action  of  the  centrif u- ' 
gal  force.     On  the  other  hand,  the  opinion  prevails  among  a  num- 
ber of  master-mechanics  that  this  rod  does  not  possess  the  required 
flexibility  sideways,  and  we  believe  that,  on  this  account,  it  is  not 
generally  adopted.     A  difference  is  also  made  by  different  designers 
in  the  distribution  of  metal  throughout  the  transverse  section ;  the  proportions  given 
in  Fig.  464  we  find   sometimes   adopted,  whereas,  for  the  same   class   and  size  of 
engines,  we  occasionally  find  the  proportions  of  the  cross-section  to  be  like  those  shown 
in  Fig.  419. 

The  type  of  side-rod  shown  in  Figs.  462,  463  has  been  used  on  some  railroads  for 
a  number  of  years,  and  is  said  to  bo  one  of  the  best  type  of  rods  in  use.  As  will  be 
seen,  it  is  made  deeper  at  the  center  than  at  the  ends,  but  its  thickness  at  the  center  is 


Fig.  464 


308  MODERN  LOCOMOTIVE   CONSTRUCTION. 

less  than  at  the  ends.  In  it  are  combined  the  best  features  of  the  rods  shown  in  Figs. 
458  and  4GO.  On  account  of  its  increased  depth  at  the  center,  it  is  stronger  to  resist 
the  action  of  the  centrifugal  force  than  the  rod  shown  in  Fig.  458 ;  and  on  account  of 
its  decreased  thickness  at  the  center,  it  has  a  greater  flexibility  than  the  one  shown  in 
Fig.  460.  But  it  is  an  expensive  rod  to  make,  and  therefore  we  believe  it  is  not  as 
often  adopted  as  its  merits  deserve. 

PROPORTIONS  OF   SIDE-RODS. 

329.  In  Art.  328  it  was  seen  that  there  are  many  causes  which  at  any  time 
may  increase  the   stress  on  a  side-rod,  and  may  increase  it  to  such  an  extent  as 
will  cause  fracture  of  or  injury  to  the  rod.     The  total  stress  to  which  the  side-rods 
may  at  any  time  be  subjected  can  only  be  determined  by  experience  and  close  obser- 
vation of  the  behavior  of  the  side-rods  in  every-day  service,  and  facts  obtained  in  this 
way  will  enable  us  to  establish  rules  for  guidance  in  designing  other  side-rods  for 
similar  locomotives.     The  following  rules  are  empirical,  and  will  hold  true  only  within 
the  limits  of  ordinary  locomotive  practice. 

In  eight-wheeled  passenger  engines,  the  side-rods  are  longer  than  those  used  in 
consolidation  engines ;  and  since  the  length  of  the  rods  has  an  important  bearing  upon 
the  size  of  the  cross-section,  that  is  to  say,  for  longer  rods  we  require  a  greater  cross- 
section,  it  follows  that  the  side-rods  for  consolidation  engines  have  generally  a  smaller 
cross-section  than  that  of  the  rods  for  passenger  engines,  diameter  of  cylinders  and 
maximum  steam  pressure  being  the  same  in  both  cases.  Again,  in  consolidation 
engines  the  front  and  rear  side-rods  have  less  work  to  do  than  the  central  side-rod, 
and  therefore  the  cross-section  of  the  latter  rod  is  generally  made  greater  than  that  of 
the  front  and  rear  rods. 

In  Mogul  and  ten-wheeled  engines,  the  length  of  the  rear  side-rods  is  generally 
equal  to  the  length  of  the  side-rods  in  passenger  engines  of  the  same  power ;  but  the 
front  side-rods  in  the  former  classes  of  engines  are  generally  shorter,  and  therefore 
the  cross-sectional  area  of  the  front  rods  is  often  made  less  than  that  of  the  rear  side- 
rods.  Yet  this  is  not  a  universal  practice,  as  in  these  classes  of  engines  we  meet  with 
many  in  which  all  the  side-rods  are  of  equal  cross-section.  Similar  remarks  apply  to 
consolidation  engines ;  that  is  to  say,  in  a  number  of  engines  of  this  class,  there  is  no 
difference  made'  in  the  cross-sectional  area  of  the  side-rods,  whereas  in  others  the 
cross-sectional  areas  for  the  front  and  rear  side-rods  are  made  less  than  that  of  the 
central  rods. 

Here,  then,  we  see  that  practice  differs,  and  therefore  in  establishing  our  rules 
we  shall  follow  the  average  of  good  practice. 

330.  In  comparing  the  side-  and  main-rods  made  by  different  builders,  we  find  that 
the  cross-sectional  area  of  the  side-rods  is  generally  about  ten  per  cent,  less  than  the 
cross-section  at  the  smallest  part  of  the  main-rod.     Consequently,  to  find  the  cross- 
section  of  side-rods  for  passenger  engines,  we  have  the  following  rule : 

RULE  68. — Divide  the  total  maximum  steam  pressure  on  the  piston  by  5,500 ;  the 
quotient  will  be  the  number  of  square  inches  in  the  cross-sectional  area  of  the  side- 
rod  for  eight-wheeled  passenger  engines. 


.M»IH:I;\  LOCOMOTIVE  CONSTRUCTION.  309 

Kx  AMPLE  103.—  In  a  passenger  engine  having  cylinders  18  inches  in  diameter,  the 
maximum  steam  pressure  on  the  piston  is  140  pounds  per  square  inch;  what  should 
he  the  cross-sectional  area  of  the  side-rod? 

The  maximum  steam  pressure  on  the  piston  is 

254.47  x  140  =  35625.8  pounds, 

35625  8 
and  '          -  =  6.47  square  inches  in  the  cross-sectional  area  of  the  side-rod. 


In  practice  a  difference  exists  between  the  ratio  of  the  thickness  to  depth  of  the 
side-rods,  but  observation  indicates  that  two  and  a  half  times  the  thickness  for  the 
depth  is  a  good  proportion  ;  and  this  we  shall  adopt  for  side-rods  in  passenger  engines. 
Hence  we  have  the  following  rule: 

RULE  69.—  To  find  the  thickness  and  depth  of  side-rods  for  passenger  engines  : 
Multiply  the  cross-sectional  area  of  the  side-rod  (as  found  by  Rule  68)  by  2,  and  divide 
the  product  by  5  ;  the  square  root  of  this  quotient  will  be  the  thickness.  Multiply  this 
thickness  by  2£  ;  the  product  will  be  the  depth  of  the  side-rod. 

EXAMPLE  104.  —  Find  the  thickness  and  depth  for  a  side-rod  for  an  eight-wheeled 
passenger  engine  with  cylinder  17  inches  diameter  ;  maximum  steam  pressure  on  the 
piston,  150  pounds  per  square  inch. 

The  maximum  steam  pressure  on  the  piston  is  226.98  x  150  =  34,047  pounds. 
According  to  Rule  68,  the  cross-sectional  area  of  the  side-rod  should  be 

34047 
_.  ,  ,r  =  6.19  square  inches. 

According  to  Rule  69,  the  thickness  of  the  side-rod  should  be 


(tin  V  V 

-  =  1.57  inches  =  1-ft  inches  nearly. 
5 

And  the  depth  should  be  1.57  x  2.5  =  3.92  inches  =  31f  inches  nearly.  Hence  the 
side-rod  should  be  1  &  inches  thick,  and  3  jf  inches  deep. 

The  foregoing  rules  will  only  hold  true  for  side-rods  made  of  the  best  quality  of 
hammered  iron,  whose  lengths  for  cylinders  9",  10",  and  11"  in  diameter  do  not  exceed 
6  feet  6  inches ;  for  12",  13",  14"  cylinders,  7  feet  6  inches ;  for  15",  16",  17"  cylinders, 
8  feet  6  inches;  and  for  cylinders  18  inches  diameter  and  upwards,  9  feet.  For 
shorter  rods  in  this  class  of  engines,  the  cross-sectional  area  could  be  somewhat 
reduced;  but  for  the  sake  of  uniformity  of  templates,  etc.,  such  close  adjustment  of 
the  cross-sectional  area  to  the  length  of  the  side-rod  is  not  usually  observed. 

The  appended  tables  have  been  arranged  by  the  foregoing  rules. 

331.  From  what  has  been  said  in  the  beginning  of  the  foregoing  article,  it  will  be  seen 
that  the  dimensions  of  the  side-roils  given  in  the  following  tables  can  be  used  for  the 
rear  side-rods  in  Mogul,  and  also  for  the  rear  side-rods  in  ten-wheeled  engines.  And  since 
the  front  side-rods  in  these  classes  of  engines  are  shorter  than  the  rear  ones,  we  believe 
it  to  be  good  practice  to  reduce  the  cross-sectional  area  of  the  front  roils.  In  nearly 
all  Mogul  and  ten-wheeled  engines,  the  thickness  of  the  roar  side-rods  is  equal  to  the 
thickness  of  the  front  ones;  consequently,  when  we  wish  to  reduce  the  cross-sectional 


310 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


area  of  the  front  rods,  we  have  only  to  reduce  their  depth,  according  to  the  following 
rule. 

EULE  70. — To  find  the  depth  for  the  front  side-rod :  Multiply  the  depth  of  the 
rear  side-rod  (taken  from  the  following  tables)  by  .08 ;  and  subtract  the  product  from 
the  depth  of  the  rear  side-rod ;  the  remainder  will  be  the  depth  of  the  front  rod. 

EXAMPLE  105. — What  should  be  the  dimensions  of  the  cross-section  for  a  front 
side-rod  in  a  Mogul  engine,  having  cylinders  18  inches  diameter,  maximum  steam 
pressure  on  the  piston  140  pounds  per  square  inch? 

Looking  in  Table  29,  we  find  that  for  a  cylinder  18  inches  diameter,  the  rear  side- 
rod  should  be  If  inches  thick,  and  4  inches  deep.  Since  the  thickness  of  the  rods  is 
not  changed,  we  know  that  the  thickness  of  the  front  side-rod  should  be  If  inches. 
According  to  Rule  70,  the  depth  of  the  front  rod  should  be  equal  to  4  inches 
-  (4  x  .08)  =  3.68,  say  3f  inches. 

TABLE   27. 

THICKNESS  AND  DEPTH  OP  SIDE-RODS  OP  UNIFORM  RECTANGULAR  CROSS-SECTION,  MADE  OP  THE 
BEST  HAMMERED  IRON,  FOR  EIGHT-WHEELED  PASSENGER  ENGINES.  MAXIMUM  STEAM  PRESSURE 
ON  THE  PISTON,  120  POUNDS  PER  SQUARE  INCH. 


Diameter  of  Cylinders. 

Thickness. 

Depth. 

9" 

i" 

11" 

10" 

Jr; 

2ft" 

11" 

2J" 

12" 

i" 

2ft" 

13" 

ift" 

-'!  i1," 

14" 

ift" 

21" 

15" 

H" 

34" 

16" 

ift;; 

3ft" 

17" 

3i" 

18" 

14" 

3fi" 

19" 

ift" 

W' 

20" 

if 

44" 

21" 

ir 

4ft" 

22" 

ifl" 

4i" 

TABLE    28. 

THICKNESS  AND  DEPTH  OP  SIDE-RODS  OF  UNIFORM  RECTANGULAR  CROSS-SECTION,  MADE  OP  THE 
BEST  HAMMERED  IRON,  FOR  EIGHT-WHEELED  PASSENGER  ENGINES.  MAXIMUM  STEAM  PRESSURE 
ON  THE  PISTON,  130  POUNDS  PER  SQUARE  INCH. 


Diameter  of  Cylinders. 

Thickness. 

Depth. 

9" 

H" 

itt" 

10" 

4" 

2i" 

11" 

«" 

2f" 

12" 

i" 

2ft" 

13" 

H" 

2f" 

14" 

1ft" 

3" 

15" 

U" 

3ft" 

16" 

if" 

3ft" 

17" 

W 

3|" 

18" 

ift" 

31" 

19" 

it" 

4ft" 

20" 

if 

4ft" 

21" 

Hi" 

*r 

22" 

If' 

4f" 

MODERy  LOCOXOTirE   COSSTRVCTION. 


311 


TABLE    29. 

THICKNESS  AXD  DEPTH  OF  SIDE-RODS  OP  UNIFORM  RECTANGULAR  CROSS-SECTION,  MADE  OF  THE 
HEST  HAMMKRED  IRON,  FOR  EIGHT-WHEELED  PASSENGER  ENGINES.  MAXIMUM  STEAM  PRESSURE 
ON  THE  PISTON,  140  POUNDS  PER  SQUARE  INCH. 


Diameter  of  Cylinders. 

Thickness. 

Depth. 

9" 

W 

2" 

10" 

1" 

2ft" 

11" 

+4" 

2ft" 

12" 

1ft" 

2|" 

13" 

H" 

2J" 

14" 

U" 

3ft" 

15" 

1ft" 

3ft" 

16" 

17" 

1ft" 
H" 

3ft" 
3f" 

18" 
19" 
20" 
21" 
22" 

It" 
Itt" 
If 

li" 
2" 

4" 
4J" 

i 

TABLE    30. 

THICKNESS  AND  DEPTH  OF  SIDE-RODS  OF  UNIFORM  RECTANGULAR  CROSS-SECTION,  MADE  OF  THE 
BEST  HAMMERED  IRON,  FOR  EIGHT-WHEELED  PASSENGER  ENGINES.  MAXIMUM  STEAM  PRESSURE 
ON  THE  PISTON,  150  POUNDS  PER  SQUARE  INCH. 


Diameter  of  Cylinder-. 

Thickness. 

Depth. 

9" 

48" 

2ft" 

10" 

1  ft// 

16 

2ft" 

11" 

12" 

u;; 

2t" 

13" 

3" 

14" 

ift" 

3i" 

15" 

U" 

Q  7    // 

16" 

u",, 

3li" 

17" 

3)f" 

18" 

it" 

44" 

19" 
20" 
21" 

22" 

it" 

V 

2" 

1" 

TABLE  31. 

THICKNESS  AND  DEPTH  OF  SIDE-RODS  OF  UNIFORM  RECTANGULAR  CROSS-SECTION,  MADE  OF  THE 
UKST  HA.M.MKKED  IRON",  FOR  EIGHT-WHEELED  PASSENGER  ENGINES.  5IAXIMUSI  STEAM  PRESSURE 
ON  THE  PISTON,  160  POUNDS  PER  SQUARE  INCH. 


Diameter  of  Cylinders. 

Thickness. 

Depth. 

9" 
10" 

ft" 

2f" 

11" 

ift" 

24" 

12" 

H" 

2|" 

13' 

if 

3ft; 

14" 

1h.ii 
Te 

15" 

•  !  ', 

16" 

ir 

3{'' 

17" 

it" 

4ft' 

18" 
19" 

JB// 

1  1  a// 

4ft' 

20" 

j  1ft'/ 

4t" 

21" 

2" 

V 

22" 

2ft" 

312  MODERN  LOCOMOTIVE   CONSTRUCTION. 

332.  In  finding  the  dimensions  for  the  side-rods  for  consolidation  engines,  the  best 
mode  of  procedure  will  be  to  find  first  the  dimensions  for  the  central  one.  We  have 
already  seen  that  these  rods  are  shorter  than  those  in  passenger  engines,  and  therefore 
the  cross-sectional  area  of  the  central  side-rods  for  consolidation  engines  is  generally 
less  than  that  of  the  side-rods  for  passenger  engines.  Observation  indicates  that  the 
average  of  good  practice  is  to  give  one  square  inch  in  the  cross-sectional  area  for  every 
6,000  pounds  of  the  maximum  steam  pressure  on  the  piston  ;  and  that  the  depth  of  the 
central  side-rod  is  about  2£  times  its  thickness.  Hence  we  have  the  following  rules  : 

RULE  71.  —  To  find  the  cross-sectional  area  of  the  central  side-rods  for  consolida- 
tion engines  :  Divide  the  total  maximum  steam  pressure  on  the  piston  by  6,000  ;  the 
quotient  will  be  the  number  of  square  inches  in  the  cross-sectional  area. 

EULE  72.  —  To  find  the  thickness  and  depth  of  the  central  side-rods  for  consolida- 
tion engines  :  Multiply  the  cross-sectional  area  (found  by  Eule  71)  by  2,  and  divide 
the  product  by  5  ;  the  square  root  of  this  product  will  be  the  thickness.  Multiply  this 
thickness  by  2j  ;  the  product  will  be  the  depth. 

EXAMPLE  106.  —  What  should  be  the  thickness  and  depth  of  a  central  side-rod  for 
a  consolidation  engine  having  cylinders  20  inches  diameter  ;  maximum  steam  pressure 
on  the  piston,  140  pounds  per  square  inch  I 

The  maximum  steam  pressure  on  the  piston  is  314.16  x  140  =  43982.4  pounds. 

According  to  Eule  71,  the  cross-sectional  area  should  be 

43982.4 

=  7-33  square  inches. 


And  according  to  Eule  72,  the  thickness  should  be 


7.33  x  2       .,  ,.-,  .     T 

-  =  1.71  inches. 
5 

The  depth  should  be 

1.71  x  2.5  =  4.275  inches. 

Avoiding  thirty-seconds  of  an  inch,  we  find  that  the  thickness  should  be  If  inches, 
and  the  depth,  4J  inches. 

The  front  and  rear  side-rods  in  consolidation  engines  have  less  work  to  do  than 
the  central  ones,  hence  the  cross-sectional  area  of  the  former  can  be  less  than  that  of 
the  latter.  In  nearly  all  consolidation  engines  the  thickness  of  all  the  side-rods 
remains  the  same ;  we  have  therefore  only  to  find  the  depth  of  the  front  and  rear  rods, 
which  can  at  once  be  obtained  by  deducting  a  certain  amount  from  the  depth  of  the 
central  one ;  the  remainder  will  be  the  required  depth  for  the  front  and  rear  side-rods. 

EULE  73. — To  find  the  depth  of  the  front  and  rear  side-rods  for  consolidation 
engines :  Multiply  the  depth  of  the  central  side-rod  by  .08,  and  subtract  the  product 
from  the  depth  of  the  central  one ;  the  remainder  will  be  the  depth  of  the  front  and 
rear  side-rods. 

EXAMPLE  107. — What  should  be  the  thickness  and  depth  of  all  the  side-rods  for  a 
consolidation  engine,  having  cylinders  22  inches  diameter ;  maximum  steam  pressure 
on  the  piston,  140  pounds?  We  first  find  the  dimensions  of  the  central  rod. 

The  total  maximum  steam  pressure  on  the  piston  is 

380.13  x  140  =  53218.2  pounds. 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


313 


According  to  Rule  71,  the  cross-sectional  area  of  this  rod  should  be 

53218.2 


liINN) 


=  8.86  square  inches. 


According  to  Rule  72,  the  thickness  of  all  the  side-rods  should  be 


5.86  x  2 


=  1.88  inches. 


The  depth  of  the  central  side-rod  should  be  1.88  x  2.5  =  4.70  inches.  And  the 
depth  of  the  front  and  rear  side-rods  should  be  4.70  -  (4.70  x  .08)  =  4.33  inches. 

Avoiding  thirty-seconds  of  an  inch,  we  find  that  the  thickness  of  all  the  side-rods 
should  be  1$  inches ;  the  depth  of  the  central  one  should  be  4|J  inches ;  and  that  of 
the  front  and  rear  rods,  4|  inches. 

The  appended  tables  have  been  computed  by  the  foregoing  rules. 


TABLE  32. 

THICKNESS  AND  DEPTH  OP  SIDE-RODS  OP  UNIFORM  RECTANGULAR  CROSS-SECTION,  MADE  OP  THE 
BEST  HAMMERED  IRON,  FOR  CONSOLIDATION  ENGINES.  MAXIMUM  STEAM  PRESSURE  ON  THE 
PISTON,  120  POUNDS  PER  SQUARE  INCH. 


Diameter  of  Cylinders. 

Thickness  of  all  the  Side- 
rode. 

Depth  of  Central  Side- 
rod. 

Depth  of  Front  and  Rear 
Side-rode. 

14" 

1ft" 

2f" 

21" 

15" 

i,V 

3" 

2f" 

16" 

1*" 

34" 

2J" 

17" 

1ft* 

3|" 

3iV' 

18" 

w' 

3ft" 

si* 

19" 

11" 

3f" 

3ft" 

20" 

1ft" 

:!!;:• 

3|" 

21" 

it!" 

44" 

3{|" 

22" 

it" 

4f" 

4" 

TABLE  33. 

THICKNESS  AND  DEPTH  OP  SIDE-RODS  OP  UNIFORM  RECTANGULAR  CROSS-SECTION,  MADE  OP  THE 
BEST  HAMMERKI)  IKON,  FOR  CONSOLIDATION  KM  JINKS.  MAXIMUM  STEAM  I'KKSSURE  ON  THE 
PISTON,  130  POUNDS  PER  SQUARE  INCH. 


Diameter  of  Cylinder*. 

Thickncm  of  all  the  Slde- 
rods. 

Mc'jith  of  Central  Side- 
rod. 

Depth  of  Front  and  Rear 
Side-rods. 

14" 
15" 

If 

li- 

24" 

•v 

2|" 
2(2" 

16" 

lt" 

3t" 

3ft" 

17" 

i,V 

31" 

3i" 

18" 

11" 

3»" 

3»" 

19" 

l,v 

3J" 

W' 

20" 

11" 

4ft" 

3t" 

21" 
22" 

Ifr 

4t" 
41" 

4" 
41" 

314 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


TABLE   34. 

THICKNESS  AND  DEPTH  OP  SIDE-RODS  OP  UNIFORM  RECTANGULAR  CROSS-SECTION,  MADE  OP  THE 
BEST  HAMMERED  IRON,  FOR  CONSOLIDATION  ENGINES.  MAXIMUM  STEAM  PRESSURE  ON  THE 
PISTON,  140  POUNDS  PER  SQUARE  INCH. 


Diameter  of  Cylinders'. 

Thickness  of  all  the  Side- 
rode. 

Depth  of  Central  Side- 
rod. 

Depth  of  Front  and  Rear 
Side-rods. 

14" 

1ft" 

3" 

2f" 

15" 

Ii" 

3ft" 

2ir 

16" 

If" 

3,V 

3*" 

17" 

iV 

3|" 

3f" 

18" 

H" 

31" 

3i" 

19" 

20" 

If1 

It" 

4ft" 
4i" 

3J" 
M" 

21" 

22" 

1«" 

if 

if 

4H" 

44" 
4|" 

TABLE  35. 

THICKNESS  AND  DEPTH  OF  SIDE-RODS  OF  UNIFORM  RECTANGULAR  CROSS-SECTION,  MADE  OF  THE 
BEST  HAMMERED  IRON,  FOR  CONSOLIDATION  ENGINES.  MAXIMUM  STEAM  PRESSURE  ON  THE 
PISTON,  150  POUNDS  PER  SQUARE  INCH. 


Diameter  of  Cylinders. 

Thickness  of  all  the  Side- 
rods, 

Depth  of  Central  Side- 
rod. 

Depth  of  Front  and  Rear 
Side-rods. 

14" 
15" 

14" 

ift" 

3ft" 
8ft" 

ati" 

3ft" 

16" 

W' 

31" 

3i" 

17" 

ii" 

3t" 

3ft" 

18" 

l*" 

4" 

3|" 

19" 

itt" 

4ft" 

3|" 

20" 

H" 

4ft" 

4ft" 

21" 

if" 

4|" 

& 

22" 

ill" 

41" 

4ft" 

TABLE  36. 

THICKNESS  AND  DEPTH  OF  SIDE-RODS  OF  UNIFORM  RECTANGULAR  CROSS-SECTION,  MADE  OF  THE 
BEST  HAMMERED  IRON,  FOR  CONSOLIDATION  ENGINES.  MAXIMUM  STEAM  PRESSURE  ON  THE 
PISTON,  ICO  POUNDS  PER  SQUARE  INCH. 


Diameter  of  Cylinders. 

Thickness  of  all  the  Side- 
rods. 

Depth  of  Central  Side- 
rod. 

Depth  of  Front  and  Rear 
Side-rods. 

14" 

Ii" 

3ft" 

art" 

15" 

If" 

3ft" 

34" 

16" 

1ft" 

3|" 

3f" 

17" 

1ft" 

31" 

3ft" 

18" 

If 

4i" 

3f" 

19" 

If" 

4ft" 

4" 

20" 

ill" 

4ft" 

4ft" 

21" 

l«" 

4||" 

4,y 

22" 

2" 

5" 

4f" 

MODEItX  LOCOMOTIVE    CONSTRUCTION. 


315 


ROD   BRASSES. 

333.  All  main-  and  side-rods  for  which  straps  are  used  are  provided  with  brass 
boxes,  generally  called  "  brasses."  These  are  made  in  pairs,  one  brass  being  an  exact 
duplicate  of  the  other. 

All  main-rod  brasses  for  the  crank-pin  should  be  babbitted.  We  have  known  a 
few  instances  in  which  no  Babbitt  metal  was  used,  but  such,  as  far  as  we  have 
seen,  proved  to  be  a  failure.  Even  when  these  boxes  were  made  of  phosphor  bronze, 
Babbitt  metal  was  required  to  keep  them  cool. 

Figs.  465,  466  represent  the  brasses  in  a  main-rod  for  the  crank-pin,  and  show  the 


rr 


"I 


i  i  * —  '  i!  y  I j 

F.n,.465        L-'  1TTT 

JLU 


H -S« H 

Fig. "466 


i J 


amount  and  the  location  of  the  Babbitt  metal  a  a  «2  a.2  in  these  brasses.  Sometimes 
these  strips  of  Babbitt  metal  are  made  of  uniform  width  throughout,  but  we  believe 
the  best  practice  is  to  make  the  ends  nar- 
rower than  the  central  parts,  as  shown  in 
Fig.  466,  which  will  prevent  the  babbitt 
from  slipping  out  of  position. 

In  many  cases  these  strips  are  placed 
at  equal  distances  apart,  as  shown  in  Fig. 
465,  but  we  believe  the  best  practice  is  to 
increase  the  distances  between  the  strips 
jind  the  joint  of  the  brasses,  so  as  to  bring 
the  center  of  the  strips  on  lines  drawn  from 
the  center  of  the  hole  to  the  corners  of 
the  brasses,  or  nearly  so.  With  this  ar- 
rangement the  strips  will  be  in  those  portions  of  the  brasses  which  contain  the 
greatest  amount  of  metal,  and  will  not  be  so  detrimental  to  the  strength  of  the  brasses 
as  in  the  positions  shown. 


Fig.  467- 


Fig.  468 


316 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


Side-rod  brasses  should  also  be  babbitted,  but  there  are  many  in  use  without  it. 
When  Babbitt  metal  is  used,  it  is  inserted  either  as  shown  in  Figs.  465,  466,  or  as 
shown  in  Figs.  467,  468.  The  former  method  we  believe  to  be  the  best  one,  because, 
with  the  Babbitt  metal  extending  clear  across  the  brasses,  the  crank-pin  will  wear 
more  evenly  than  with  the  Babbitt  metal  inserted  as  shown  in  Figs.  467,  468.  This 
latter  method  should  never  be  adopted  for  the  crank-pin  brasses  in  main-rods. 

334.  In  many  locomotives  the  side-rod  brasses  are  made  similar  in  form  to  that 
shown  in  Figs.  467,  468,  and  in  a  few  instances  they  are  made  similar  to  that  shown  in 


i'j  i  fon  i  1  H^j    *"*rf*l 

°--,—fu  .Fty.  470 

i  i  *f 


.i    I  UL.JJ  i 

inJ  I     j  L 

h---,— I  L- !— TTI 

Lf»4<J     JF%.  4<*9       '-<U-<U 

1 W-J 

Figs.  469,  470.  The  difference  between  these  two  forms  is,  that  one  set  is  provided 
with  caps  e,  as  shown  in  Fig.  470,  which  are  cast  to  the  brasses ;  tho  other  set  is  made 
up  of  plain  brasses — that  is  to  say,  they  have  no  caps. 

These  caps  cover  the  end  of  the  crank-pin ;  their  purpose  is  to  keep  the  crank-pin, 
as  much  as  possible,  free  from  dust.  The  caps  answer  the  purpose  for  which  they 
have  been  designed ;  but  they  are  detrimental  to  the  examination  of  the  condition  of 
the  pin  which  they  cover ;  and  consequently  they  are  not  as  frequently  adopted  as  the 
ordinary  plain  brasses. 

In  Fig.  465  it  will  be  noticed  that  the  flanges  of  the  brasses  do  not  extend  to  the 
edges  of  the  strap ;  there  are  many  main-rods  as  well  as  side-rods  which  have  brasses 
of  this  design.  But  there  is  an  objection  against  the  flanges  stopping  short  of  the 
edges  of  the  strap.  The  brasses  are  exposed  to  considerable  dust,  and  as  soon  as  they 
become  a  little  loose,  the  dust  will  work  in  between  the  flanges  and  strap,  and  wear 
ridges  in  the  latter,  so  that,  when  it  becomes  necessary  to  replace  the  brasses,  the 
straps  must  be  re-planed  before  the  new  brasses  can  be  used.  Re-planing  straps  should 
be  avoided  as  much  as  possible,  as  it  reduces  their  strength;  consequently  we 
must  prevent,  as  much  as  possible,  unequal  wear  and  the  formation  of  ridges  in  the 
sides  of  the  straps.  This  desired  result  is  obtained,  to  some  extent,  by  allowing  the 


MODERN  LOCOMOTITE   CONSTRUCTION. 


317 


flanges  of  the  brasses  to  cover  the  whole  width  i  (Fig.  469)  of  the  strap,  and  also  cover 
the  solid  end  of  the  same.  Since  one  brass  is  a  duplicate  of  the  other  one,  the  depth 
of  the  flange  at  C  appears  to  be  and  is  excessive ;  yet,  on  account  of  the  practical 
advantages  gained  by  making  the  flanges  so  deep,  there  are,  probably,  at  present  more 
brasses  with  deep  flanges  used  than  brasses  with  flanges  stopping  short  of  edges  of 
the  strap. 

335.  The  thickness  I  (Fig.  469)  of  the  metal  at  the  joint  of  the  brasses,  in  main- 
and  side-rods,  is  generally  about  J  of  an  inch.  . 

The  thickness  k  of  metal  between  the  pin  and  the  butt  end  of  the  rods,  and  also 
the  thickness  of  the  flanges,  are  given  in  the  following  table : 

THICKNESS   k    (FIG.   469)   OP  METAL  IN  MAIN-  AND  SIDE-ROD  BRASSES. 


Diameter  of  Cylinders. 

Thickness  of  Metal  at  *, 
Fig.  48V. 

Thickness  of  Flanges. 

9" 

4" 

r 

10" 

t" 

i" 

11" 

1" 

•ft" 

12" 

J" 

»" 

13" 

i" 

i" 

14" 

i" 

v 

15" 
16" 

*" 

i" 

i 

17" 

l" 

»" 

18" 

l" 

*" 

19" 
20" 

1" 
l" 

$ 

21" 

H" 

r 

22" 

H" 

\" 

The  length  of  side-rods  should  be  ascertained  by  actual  measurement  when  the 
engine  is  hot.  The  brasses  should  have  a  somewhat  loose  fit  on  the  crank-pins,  so 
that,  when  the  engine  is  in  working  order  under  steam,  the  side-rods  can  be  moved  on 
the  crank-pin  to  just  a  perceptible  extent. 

The  following  proportions  of  the  different  metals  for  main-  and  side-rod  brasses 
we  believe  will  give  good  satisfaction : 

Six  pounds  of  copper,  one  pound  of  tin ;  to  one  hundred  pounds  of  this  mixture 
add  one-half  pound  of  zinc  and  one-half  pound  of  lead. 


CRANK-PINS. 

336.  Crank-pins  are  made  either  of  steel  or  of  the  best  quality  of  hammered  iron. 
In  order  to  reduce  the  wear  on  iron  crank-pins,  they  are  frequently  case-hardened. 
During  the  process  of  case-hardening,  the  crank-pin  will,  to  some  extent,  alter  its  form, 
and  therefore,  after  case-hardening,  it  must  be  trued  up.  In  so  doing,  the  case-hard- 
ened surface  may  be  reduced  to  an  uneven  thickness,  which  in  time  may  produce  an 
uneven  wear,  and  interfere  with  the  cool  and  smooth  running  of  the  engine.  On  the 
other  hand,  steel  crank-pins  do  not  need  to  be  case-hardened,  they  wear  well  without 
it,  and  therefore  the  chances  of  obtaining  a  wearing  surface  of  different  degrees  of 
hardness  are  lessened,  and  the  causes  of  heating  and  uneven  wear  are  to  some  extent 
removed.  Yet  a  wrought-iron  pin  has  an  advantage  over  a  steel  one;  the  latter,  when 


318  MODERN  LOCOMOTIVE   CONSTRUCTION. 

subjected  to  excessive  pressure — which  may  happen  even  in  the  best  designed 
engines — may  break  or  snap  off  suddenly,  and  thereby  cause  considerable  damage ;  on 
the  other  hand,  a  wrought-iron  pin  will  bend  to  a  greater  extent  before  it  breaks  than 
a  steel  one,  and  consequently  may  give,  in  many  instances,  a  timely  warning  of  an 
excessive  pressure,  so  that  repairs  or  changes  can  be  made  before  much  damage  has 
been  done.  But  steel  pins  can  resist  a  greater  pressure  than  iron  ones,  and,  since  they 
do  not  need  to  be  case-hardened,  they  are  often  preferred.  Hence  on  some  roads  we 
find  steel  pins  used  exclusively,  and  on  other  roads  iron  pins  are  adopted.  We  prefer 
steel  pins. 

337.  From  the  foregoing  remarks  we  infer  that  in  designing  a  crank-pin  we  must 
keep  in  view  its  strength,  and  also  its  liability  of  heating.     To  prevent  heating,  we 
must  have  a  sufficient  bearing  surface,  and,  when  a  sufficient  bearing  surface  has 
been   provided,   then  the  crank-pins,   having  such   proportions  as  are  adopted  in 
modern  locomotive  practice,  will  also  be  strong  enough  for  the  work  they  have  to  do. 
Hence,  in  determining  the  dimensions  of  a  crank-pin  we  shall  be  guided  mostly  by 
the  pi'essure  which  the  crank-pins  have  to  resist.     The  pressure  on  the  crank-pin  is 
estimated  by  the  pressure  on  its  projected  area ;  that  is  to  say,  by  the  pressure  on  a 
rectangular  surface,  whose  length  and  breadth  are  equal  to  the  length  and  diameter 
of  the  crank-pin  journal. 

In  comparing  the  pressure  per  square  inch  of  projected  area  of  the  crank-pins  as 
made  by  different  makers,  we  find  a  great  difference;  indeed,  in  some  instances  we 
find  the  pressure  per  square  inch  on  the  projected  area  to  be  about  1,000  pounds ;  in 
other  cases  the  pressure  is  nearly  2,000  pounds  per  square  inch.  The  low  pressures 
on  crank-pins  we  find  to  occur  mostly  in  small  engines,  and  the  higher  pressures 
mostly  in  large  engines ;  which  seems  to  indicate  that  the  crank-pins  in  a  number  of 
small  engines  are  somewhat  large,  and  in  a  number  of  large  engines  the  crank-pins  are 
too  small.  The  truth  of  these  conclusions,  we  believe,  is  confirmed  by  experience  and 
the  results  in  practice. 

In  the  rules  which  are  to  follow,  we  shall  adopt  1,600  pounds  per  square  inch  of 
projected  area  of  the  crank-pin,  and  determine  the  size  of  all  crank-pins  according  to 
this  pressure. 

CBANK-PINS   FOR  EIGHT-WHEELED   PASSENGER   ENGINES. 

338.  In  designing  a  main  crank-pin,  care  must  be  taken  not  to  make  its  journals 
too  long,  because  an  increase  of  length  will  weaken  the  pin ;  the  diameters  of  the 
journals  should  not  be  larger  than  necessary,  because  enlarging  the  diameters  will 
occasion  a  loss  of  work  due  to  friction.     Consequently,  there  should  be  a  ratio  between 
the  diameter  and  the  length  of  a  locomotive  crank-pin  journal ;  and  this  ratio  can 
best  be  established  by  the  proportions  of  crank-pins  now  in  actual  and  successful 
service. 

In  all  locomotives  which  have  more  than  one  pair  of  drivers,  the  main  crank-pins 
have  two  journals,  as  shown  in  Fig.  471.  One  of  these  journals  is  the  main-rod 
journal ;  the  other  is  the  side-rod  journal. 

In  wide-gauge  (4  feet  8£  inches)  eight-wheeled  passenger  engines,  the  main-rod  is 
nearly  always  placed  next  to  the  wheels;  in  narrow-gauge  eight- wheeled  passenger 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


319 


Gui- 
lt is 


Fig.  471 


engines  we  are  frequently  obliged  to  place  the  side-rod  next  to  the  wheel.  If  Fig.  471 
iv|>rt>sfiits  Ji  crank-pin  designed  for  an  eight-wheeled  passenger  engine  (4  feet  8£  inches 
•MI me),  then  the  journal  B  will  represent  the  main-rod  journal ;  and  the  journal  C  will 
represent  the  side-rod  journal.  The  portion  A  of  the  pin  which  is  pressed  into  the 
wheel  is  called  the  wheel  fit. 

Let  Fig.  471  represent  a  main  crank-pin  for  an  eight-wheeled  passenger  engine :  it 
is  required  to  find  the  diameter  I)  and  the  length  L  of  the  main-rod  journal  B ;  it  is 
also  required  to  find  the  diameter  d  and  the  length  I  of  the  side-rod  journal  C. 

Let  us  commence  by  finding  the  dimensions  of  the  main-rod  journal  B. 
first  step  will  be  to  establish  a  ratio  between  its  diameter  D  and  its  length  L. 
not  often  that  we  find  in  eight- 
wheeled  passenger  engines  a 
crank-pin  in  which  the  length 
of  the  main  journal  exceeds  its 
diameter ;  but  there  are  a  num- 
ber of  eight-wheeled  passenger 
engines  in  which  the  length  and 
diameter  of  the  main-rod  journal 
are  equal  to  each  other;  and 
lastly,  we  believe  it  is  safe  to  say 
that,  in  the  majority  of  engines, 

the  diameter  of  the  main  journal  is  greater  than  its  length.  In  the  latter  cases  the 
ratio  between  the  diameters  and  lengths,  as  made  by  different  builders,  varies  some- 
what ;  but  good  practice  seems  to  indicate  that  the  diameter  of  the  main-rod  journal 
should  be  about  1J  times  greater  than  its  length.  Occasionally  crank-pins  need  to  bo 
trued  up  a  little,  for  which  a  small  allowance  should  be  made.  Hence,  in  determining 
the  dimensions  of  the  main-rod  journal,  we  shall  first  find  its  diameter  and  length, 
whose  ratio  is  as  1J  to  1,  and  then  add,  to  allow  for  wear,  about  -fa  to  £  inch  to  the 
diameter  thus  found.  We  say  ab&ut  fa  to  £  inch,  because  the  diameter  determined 
by  calculation  will  in  many  cases  contain  a  fraction  of  an  inch,  and  when  this 
fraction  does  not  contain  even  &  inch,  we  shall  then  add  a  little  less,  or  in  some  cases 
a  trifle  more,  than  £  inch,  so  as  to  make  the  fraction  divisible  by  £  inch ;  indeed,  many 
locomotive  builders  do  not  adopt  a  diameter  which  cannot  be  divided  by  £  inch.  If 
the  diameter  obtained  by  calculation  contains  even  J  inch,  we  simply  add  4  inch 
for  wear.  When  the  length  of  the  journal,  as  found  by  calculation,  contains  frac- 
tions which  cannot  be  divided  by  £  inch,  we  simply  take  the  nearest  fraction 
divisible  by  it  to  that  found;  hence,  in  some  cases  the  lengths  adopted  may  be  a 
little  less,  and  in  others  a  little  greater  than  that  found  by  calculation. 

For  the  sake  of  simplicity,  we  shall  assume  that  the  whole  steam  pressure  on  the 
piston  is  exerted  to  turn  the  wheels,  allowing  nothing  for  the  friction  of  piston,  etc.; 
we  shall  also  neglect  the  pressure  due  to  the  obliquity  of  the  main-rod. 

Now,  since  the  pressure  per  square  inch  of  projected  area  of  a  steel  crank-pin 
journal  is  to  be  1,600  pounds,  we  can  readily  find,  under  the  foregoing  conditions,  this 
area  when  the  steam  pressure  on  the  piston  is  known,  thus: 

RULE  74. — For    crank-pins    made    of    steel,  divide  the    total    maximum    steam 


320  MODERN  LOCOMOTIVE  CONSTRUCTION. 

pressure  on  the  piston  by  1,600  ;  the  quotient  will  be  the  number  of  square  inches  in 
the  projected  area  of  the  main-rod  journal  for  eight-  wheeled  passenger  engines. 

Now,  since  the  diameter  of  the  journal  is  to  be  l£  times  its  length,  we  have  the 
following  rule  : 

EULE  75.  —  Multiply  the  projected  area  (as  found  by  Eule  74)  by  8,  and  divide  the 
product  by  9;  extract  the  square  root  of  the  quotient;  the  result  will  be  the 
length  of  the  journal.  Multiply  this  length  by  1.125,  and  add  for  wear  so  as  to  make 
the  diameter  divisible  by  £  inch  ;  the  sum  will  be  the  required  diameter. 

EXAMPLE  108.  —  Find  the  dimensions  for  the  main-rod  crank-pin  journal  for  a 
passenger  engine  having  cylinders  18  inches  diameter  ;  maximum  steam  pressure  per 
square  inch  of  piston,  130  pounds. 

The  area  of  an  18-inch  piston  is  254.47  square  inches  ;  hence,  the  maximum  steam 
pressure  on  the  piston  will  be 

254.47  x  130  =  33081.10  pounds. 

According  to  Eule  74,  the  number  of  square  inches  in  the  projected  area  of  the 
main-rod  journal  will  be 

33081.10 

=  20.67  square  inches. 


According  to  Eule  75,  we  find  the  length  to  be 


8  x  20.67       .  00  .     , 
— ^-    -  =  4.28  inches. 

t7 

And  the  diameter  we  find  to  be 

4.28  x  1.125  =  4.815  inches; 

adding  to  this  diameter  .06  (which  in  this  case  is  a  little  less  than  ^  inch),  we 
obtain  4J  inches.  Hence,  the  diameter  of  this  journal  should  be  4J  inches,  and  its 
length,  4J  inches. 

339.  To  determine  the  dimensions  of  the  side-rod  journal  we  proceed  in  manner 
similar  to  that  adopted  for  finding  the  dimensions  of  the  main  journal.  Our  first  step 
will  be  to  compute  the  projected  area.  In  good  practice  we  find  that  in  eight- wheeled 
passenger  engines  the  projected  area  of  the  side-rod  journal  is  equal  to  about  65  to  75 
per  cent,  of  that  of  the  main-rod  journal ;  we  believe  that  about  67  per  cent,  is  a  good 
proportion.  Hence,  when  we  know  the  projected  area  of  the  main-rod  journal,  that 
of  the  side-rod  journal  may  be  found  by  making  it  equal  to  about  67  per  cent,  of 
the  former.  But  this  result  can  be  obtained  in  a  more  direct  way  by  the  following 
rule,  which  will  give  results  agreeing  well  with  the  average  good  practice. 

EULE  76. — For  steel  crank-pins,  divide  the  maximum  pressure  on  the  piston  by 
2,400 ;  the  quotient  will  be  the  number  of  square  inches  in  the  projected  area  of  a  side- 
rod  journal  for  eight-wheeled  passenger  engines. 

In  many  locomotives  we  find  the  length  of  the  side-rod  journal  to  be  equal  to  its 
diameter ;  on  the  other  hand,  we  find  as  many,  if  not  a  greater  number  of  engines,  in 
which  the  diameter  of  the  side-rod  journal  is  greater  than  its  length.  In  the 
latter  class,  the  ratio  between  the  diameters  and  lengths,  as  made  by  different 
makers,  varies  somewhat,  but  good  practice  seems  to  indicate  that  this  ratio  should 


!.(><•<  >Morni:  i-o\sri;i  <-THI\. 


321 


be  equal  to  that  of  the  main-rod  journal,  namely,  the  diamter  should  be  l£  times 
greater  than  the  length;  and  these  porportions  we  shall  adopt. 

To  the  diameter  thus  found  we  shall  also  add  ^  to  £  inch,  so  as  to  avoid 
fractious  which  cannot  be  divided  by  &  inch.  As  for  the  length  of  the  side-rod 
journals,  we  .shall  simply  adopt  such  as  can  be  divided  by  £  inch,  and  which  will  be  the 
nearest  to  the  fraction  found  by  calculation,  and  therefore  may  in  some  cases  be  a 
little  less  and  in  others  greater  than  that  obtained  by  the  rule. 

RULE  77. — To  find  Ihe  length  and  diameter  of  a  steel  side-rod  journal  for  eight- 
wheeled  passenger  engine.  Multiply  the  projected  area  of  the  side-rod  journal  (as 
found  by  Eule  76)  by  8,  and  divide  the  product  by  9;  extract  the  square  root  of  the 
quotient ;  the  result  will  be  the  required  length.  Multiply  the  length  by  1.125,  add  to 
the  product  from  ^  to  £  inch,  so  as  to  obtain  a  numerical  value  divisible  by  &  inch ; 
the  suni  will  be  the  required  diameter. 

.  EXAMPLE  109. — Find  the  dimensions  for  the  side-rod  journal  for  a  passenger 
engine  having  cylinders  18  inches  diameter;  maximum  steam  pressure  per  square  inch 
of  piston,  130  pounds. 

We  have  already  found,  in  Example  108,  that  the  maximum  steam  pressure  on 
the  piston  is  33081.10  pounds ;  hence,  according  to  Rule  76,  we  have 

33081.10 
., ,  ,         =  13.78  square  inches, 

which  is  the  projected  area  of  the  side-rod  journal. 
According  to  Rule  77,  the  length  should  be 


v7' 


8  x  13.78 
9 


=  3.49  inches. 


For  the  diameter  we  have  3.49  x  1.125  =  3.926  inches ;  to  this  add  .199  inch ;  the 
sum  will  be  4.1 25,  which  is  the  required  diameter;  hence,  the  side-rod  journal  should 
be  4£  inches  diameter,  and  3£  inches  long. 

In  this  manner  the  diameters  and  lengths  of  steel  crank-pins  given  in  the  follow- 
ing tables  have  been  obtained. 

TABLE  37. 

PI.MEXSIOXS  OK  STEEL  CKAXK-1'IX  JOl'IiXALS  FOB  EIGHT-WHEELED  PASSEXOEU  ENGINES  (FOUE  WHEELS 
COXXEiTEPi.  MAXIMT.M  STEAM  PKESSI'KE  OX  THE  PISTON,  Il!!>  POUNDS  PER  SQUARE  INCH. 
LETTERS  AT  THE  HEAD  OF  rol.l'MXs  KEFEK  TO  FIG.  471. 


Main-rod  Journals. 

Side-rod  Journals. 

Diameter  of  Cylin- 

ders. 

Diameter 
D 

Length 

Diameter 
d 

Length 

9 

.1 

0 

If 

10 

a 

2* 

11 

11 

21 

2i 

21    . 

2 

12 

••• 

2} 

2* 

2t 

13 

3* 

3 

2* 

2* 

14 

31 

3* 

3* 

q 

15 

4 

3i 

at 

2J 

16 

4i 

w 

3* 

3 

17 

4* 

3i 

31 

34 

is 

4} 

44 

3J 

31 

19 

:. 

*f 

4* 

3* 

ta 

»i 

« 

*l 

31 

322 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


TABLE  38. 

DIMENSIONS  OF  STEEL  CRANK-PIN  JOURNALS  FOR  EIGHT-WHEELED  PASSENGER  ENGINES)  FOUR  WHEELS 
CONNECTED).  MAXIMUM  STEAM  PRESSURE  ON  THE  PISTON,  130  POUNDS  PER  SQUARE  INCH. 
LETTERS  AT  THE  HEAD  OF  COLUMNS  REFER  TO  FIG.  471. 


Dinmeter  of  Cylin- 

Main-rod Journals. 

Side-rod  Journals. 

ders. 

Diameter 
2> 

Length 
L 

Diameter 
d 

Length 

9 

84 

24 

24 

U 

10 

at 

2| 

2i 

2 

11 

3* 

2* 

24 

24 

12 

3| 

21 

2i 

2| 

13 

Bf 

34 

21 

24 

14 

31 

31 

34 

2f 

15 

4 

34 

3| 

3 

16 

41 

34 

3* 

34 

17 

*j 

4 

31 

3f 

18 

41 

4i 

44 

34 

19 

54 

44 

4i 

4 

20 

ii 

4f 

44 

31 

TABLE  39. 

DIMENSIONS  OF  STEEL  CRANK-PIN  JOURNALS  FOR  EIGHT-WHEELED  PASSENGER  ENGINES  (FOUR  WHEELS 
CONNECTED).  MAXIMUM  STEAM  PRESSURE  ON  THE  PISTON,  140  POUNDS  PER  SQUARE  INCH. 
LETTERS  AT  THE  HEAD  OF  COLUMNS  REFER  TO  FIG.  471. 


Main-rod  Journals. 

Side-rod  Journals. 

Diameter  of  Cylin- 

ders. 

Diameter 
D 

Length 

Diameter 

a 

Length 

9 

•      2f 

2i 

24 

« 

10 

21 

24 

2f 

2 

11 

34 

24 

2| 

8f 

12 

34 

3 

21 

24 

13 

34 

3± 

34 

H 

14 

4 

34 

3i 

21 

15 

4i 

3| 

34 

3 

16 

4f 

4 

34 

3i 

17 

41 

4i 

4 

34 

18 

54 

44 

4* 

31 

19 

51 

4* 

44 

31 

20 

5* 

41 

4* 

4 

TABLE  40. 

DIMENSIONS  OF  STEEL  CRANK-PIN  JOURNALS  FOR  EIGHT-WHEELED  PASSENGER  ENGINES  (FOUR  WHEELS 
CONNECTED).  MAXIMUM  STEAM  PRESSURE  ON  THE  PISTON,  150  POUNDS  PER  SQUARE  INCH. 
LETTERS  AT  THE  HEAD  OF  COLUMNS  REFER  TO  FIG.  471. 


Main-rod  Journals. 

Side-rod  Journals. 

Diameter  of  Cylin- 

ders. 

Diameter 
D 

Length 
L 

Diameter 
d 

Length 

9 

21 

2f 

ai 

11 

10 

3 

2* 

24 

24 

11 

3± 

21 

2} 

2J 

12 

Bl 

34 

3 

24 

13 

31 

3f 

34 

2* 

14 

44 

3f 

3f 

3 

15 

44 

31 

81 

3J 

16 

44 

44 

31 

3i 

17 

5 

4| 

44 

34 

18 

BJ 

*f 

4| 

3} 

19 

Bf 

41 

4f 

4 

20 

51 

54 

44 

44 

LOCOMOTIVE   COXSTRUCTIOX. 


323 


TABLE  41. 

DIMKXSFONS  OP  STEEL  CRANK-PIN  JOURNALS  FOR  EIGHT-WHEELED  PASSENGER  ENGINES  (FOUR  WHEELS 
oiNNKtTKlM.  MAXIMUM  STEAM  PRESSURE  ON  THE  PISTON,  160  POUNDS  PER  SQUARE  INCH. 
LETTERS  AT  THE  HEAD  OF  COLUMNS  REFER  TO  FIG.  471. 


Main-rod  Journals. 

Side-rod  Journals. 

T>i  m  tpr  nf  Cvlln 

(U-re. 

Diameter 
D 

Length 

Diameter 
d 

Length 

B 

2* 

2t 

24 

2 

10 

3* 

-'- 

24 

24 

11 

3f 

-: 

2* 

2t 

12 

3f 

34 

3 

2* 

13 

4 

34 

34 

2J 

H 

44 

3J 

34 

3 

15 

44 

4 

3* 

34 

16 

4| 

•H 

4 

34 

17 

54 

*i 

44 

3f 

18 

Bj 

4J 

44 

34 

19 

H 

5 

4f 

44 

20 

64 

5i 

5 

4| 

340.  Fig.  472  represents  a  rear  crank-pin  for  an  eight-wheeled  passenger  engine ; 
in  fact,  this  figure  and  Fig.  471  represent  a  pair  of  crank-pins  for  an  engine  of  this 
class. 

Figs.  474,  475  represent  another  pair  of  crank-pins  for  the  same  class  of  engine ; 
they  were  made  of  steel,  and  designed  for  an  engine  having  cylinders  17  inches  diame- 
ter. Fig.  474  represents  the  main,  and  Fig.  475,  the  rear  crank-pin ;  their  forms,  as 
will  be  seen,  differ  somewhat  from  those  of  the  pins  previously  referred  to. 

The  general  practice  is  to  make  the  diameter  of  the  outer  collar  (see  Fig.  471) 
from  1  to  l£  inches  greater  than  the  diameter  d  of  the  journal  C;  the  diameters  of  the 
middle  collar  and  the  collar  next  to  the  hub  of  the  wheel  are  generally  made  1J  inches 
greater  than  the  diameter  D  of  the  journal  7?.  The  thickness  h  of  the  outer  collar  is 
generally  J  inch,  sometimes  f  inch.  The  thickness  g  of  the  middle  collar,  and  the 
thickness/  of  the  collar  next  to  the  hub  of  the  wheel,  will  depend  upon  the  distance 
between  the  line  coinciding  with  the  axis  of  the  cylinder  and  the  face  of  the  hub  of 
wheel.  In  many  eight-wheeled  passenger  engines  the  thickness  g  is  J  inch,  and  the 
thickness/  3  inch. 

Sometimes  the  hub  of  the  wheel  is  made  exceedingly  deep,  leaving  no  room  for  a 
collar  next  to  the  hub;  in  such  cases  the  main  crank-pin  will  appear  as  shown  in 
Fig.  474.  The  diameter  of  the  wheel  fit  of  this  crank-pin  is  made  comparatively  great, 
so  as  to  obtain  a  shoulder  against  which  the  brasses  of  the  main-rod  can  bear;  this 
shoulder  generally  projects  ^  of  an  inch  beyond  the  face  of  the  hub. 

The  junctions  /  /  <>»'  the  journal  and  the  face  of  the  collars  (Fig.  471)  should  in 
every  instance  be  curved  surfaces  turned  to  a  radius  of  |  inch  for  small  crank-pins, 
and  increased  up  to  a  radius  of  J  inch  for  large  pins. 

For  many  rear  crank-pins  the  shank  E  is  formed  so  as  to  leave  a  collar  next  to  the 
hub  of  the  wheel,  as  shown  in  Fi^.  47±  The  diameter  n  of  this  collar  is  made  equal  to 
that  of  the  corresponding  one  on  the  main  crank-pin,  and  its  thickness  p  is  generally 
|  or  ?  inch.  Tin-  iliann-tfi-  ///  of  the  shank  E  is  generally  made  from  \  to  J  inch 
greater  than  the  diameter  D  of  the  main-rod  journal. 


324 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


Fig.  473  also  represents  a  rear  crank-pin  for  an  eight-wheeled  passenger  engine. 
The  shank  E  of  this  pin  has  a  uniform  taper  nearly  throughout  its  whole  length ;  the 
diameter  n  is  made— as  the  diameter  n  of  the  collar  of  the  shank  in  Fig.  472 — equal  to 
the  diameter  of  the  collar  on  the  main  crank-pin ;  that  is,  the  collar  next  to  the  hub. 


Making  the  form  of  the  shank  like  that  shown  in  Fig.  473  increases  the  weight  of  the 
pin  unnecessarily ;  it  does  no  good,  and  therefore  this  form  of  shank  is  not  recom- 
mended. Indeed,  in  order  to  reduce  the  weight  of  the  rear  crank-pin,  some  locomotive 
builders  make  its  shank  E  like  that  shown  in  Fig.  479 ;  this  rear  crank-pin  has  a  collar 
on  each  side  of  the  journal,  and  the  diameter  r,  next  to  the  inner  collar,  is  generally 
made  |  of  an  inch  greater  than  the  diameter  d  of  the  journal. 

Figs.  476,  477  also  represent  a  pair  of  crank-pins  for  an  eight-wheeled  passenger 


MODERX  LOCO.MOTITK   CONSTRUCTION. 


325 


having  cylinders  17  inches  diameter.  It  will  be  noticed  that  the  shank  E  of 
the  rear  crank-pin,  Fig.  47(5,  is  different  in  form  from  those  previously  shown.  Here 
the  diameter  s,  in  the  center  of  the  shank,  is  made  less  than  the  diameter  k,  or  the 
diameter  of  the  wheel  fit.  The  object  of  this  design  of  crank-pin  is  to  reduce  its 
rigidity,  so  that  when  the  crank-pin  is  subjected  to  a  sudden  stress  or  shock,  the 
slight  flexibility  which  it  may  possess  will  lessen  the  effect  of  the  shock,  and 


thus  reduce  the  chances  of  breaking.  Of  course,  the  proportions  of  this  crank-pin 
must  be  such  that  none  of  its  fibers  will  be  strained  beyond  the  limits  of  elasticity. 
These  crank-pins  are  used  to  quite  an  extent  on  one  of  our  prominent  roads,  and  give 
good  satisfaction.  On  other  roads  this  form  of  crank-pin  is  seldom  found.  The  reason 
why  this  form  is  not  generally  adopted  is  probably  on  account  of  the  difficulty  of 
determining  its  correct  proportions ;  in  fact,  these  proportions  are  generally  obtained 
by  a  tentative  process  rather  than  by  computation, 

Figs.  478,  479  represent  a  pair  of  crank-pins  for  an  eight-wheeled  passenger 
engine  with  cylinders  17  inches  diameter ;  they  are  also  often  used  for  the  same  class 
of  engines  with  cylinders  18  inches  diameter. 

These  crank-pins  are  designed  for  engines  which  have  solid-ended  side-rods.  One 
of  these  rods  is  shown  in  Fig.  41!). 

,'!41.  In  eight-wheeled  passenger  engines  the  diameter  of  the  wheel  fit  of  the  rear 
crank-pin  is  always  made  equal  to  that  of  the  main  crank-pin.  This  diameter  should 
never  be  made  less  than  the  diameter  I)  uf  the  journal  next  to  the  wheel;  in  fact,  we 
believe  it  to  be  good  practice  to  make  the  diameter  of  the  wheel  fit  from  £  to  4  inch 
greater  than  the  diameter  I). 


326 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


When  the  crank-pin  has  a  collar  next  to  the  wheel,  the  jtmction  of  the  wheel  fit 
and  the  face  of  the  collar  should  never  be  a  sharp  corner,  but  should  be  well  rounded  out. 

The  crank-pins  are  generally  pressed  into  the  wheels.  When  the  crank-pin 
hole  is  perfectly  true  and  smooth,  the  pin  should  be  pressed  in  with  a  press- 


F ig.  480 


ure  equal  to  about  six  tons  for  every  inch  in  diameter  of  the  wheel  fit.  When  the 
hole  is  not  perfectly  true,  which  may  be  the  result  of  shrinking  the  tire  on  the 
wheel  center  after  the  hole  for  the  crank-pin  has  been  bored,  or  if  the  hole  is  not 
perfectly  smooth,  the  pressure  may  have  to  be  increased  to  nine  tons  f or  every  inch  in 
diameter  of  the  wheel  fit.  From  these  remarks  it  appears  that  it  is  always  best  to 
shrink  the  tires  on  the  wheel  centers  before  the  holes  for  the  crank-pins  are  bored. 

342.  Fig.  480  represents  another  form  of  crank-pin  used  for  solid-ended  side-rods ; 
it  differs  from  that  shown  in  Fig.  479  in  the  fact  that  it  has  no  loose  collar  for  holding 
the  side-rod  in  position  after  the  latter  has  been  slipped  on  to  the  journal ;  instead  of  a 
loose  collar  a  groove  a  is  turned  into  the  journal  near  the  end.  The  design  of  side-rod 
used  for  this  pin  is  shown  in  Figs.  481  ,and  482.  A  solid  brass  bushing  is  forced  into 


MODERN   LOCOMOTIVE    CONSTRUCTION. 


327 


the  end  of  the  side-rod;  a  brass  plate  B  made  in  two  pieces  is  made  to  fit  into  the 
groove  a  in  the  crank-pin  journal ;  a  brass  cap  C  covers  the  end  of  the  pin ;  four  bolts 
I)  D  extend  through  the  whole  width  of  this  side-rod  end  and  hold  the  cap  C  and  the 


W  ~, 

...; 

L 

"v'5 

1 

1 

i 

plate  B  firmly  in  position,  and  also  provont  the  brass  bushing  from  turning  around 
should  it  become  loose  through  constant  service.  The  bolt-heads  are  countersunk  into 
the  flange  of  the  bushing,  as  shown  in  Fig.  483. 

Fig.  483  represents  in  detail  the  brass  bushing  A ;  Fig.  484  represents  the  brass 
plate  B;  and  Fig.  485  represents  the  brass  cap  C;  these  require  no  further  explanation. 


328 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


MAIN   CRANK-PINS   FOR  MOGUL,   TEN-WHEELED,   AND   CONSOLIDATION   ENGINES. 

343.  In  ten-wheeled,  Mogul,  and  consolidation  engines,  the  arrangement  of  the 
side-  and  main-rods  generally  differs  from  the  arrangement  of  the  rods  in  passenger 
engines.  In  the  latter  class  of  engines  we  have  seen  that  the  main-rod  takes  hold  of 
the  inner  journal  of  the  main  crank-pin ;  in  the  former  classes  of  engines  the  main-rod 

generally  takes  hold  of  the  outer  journal  of 
the  main  crank-pin,  thus  bringing  the  side- 
rods  next  to  the  wheels. 

Fig.  487  represents  a  main  crank-pin  for 
a  consolidation  engine  having  cylinders  20 
inches  diameter ;  and  Fig.  486  represents  one 
of  the  side-rod  pins  for  the  same  engine.  The 
main  crank-pins  for  consolidation,  Mogul, 
and  ten-wheeled  engines  are  generally  simi- 
lar in  form ;  hence  the  rules  by  which  the 
dimensions  of  the  crank-pin  journals  for  one  of  these  classes  of  engines  are  deter- 
mined can  also  be  used  for  finding  the  dimensions  of  similar  crank-pin  journals  for 
the  other  two  classes  of  engines. 

The  following  rules  are  for  steel  pins.  In  establishing  these  rules  we  shall  again 
be  guided  by  the  dimensions  of  the  crank-pins  at  present  in  use,  and  which  give  good 
satisfaction.  Let  us  commence  with  the  main-rod  journal  on  the  main  crank-pin ;  this 
joiirnal  will  be  the  outer  one  in  Fig.  487 ;  it  is  required  to  find  the  diameter  D  and  the 


Fig.  486 


Fig.  487 


length  L.  In  Art.  337  it  has  been  stated  that,  to  prevent  a  locomotive  crank-pin  from 
heating,  it  must  have  a  sufficient  bearing  surface,  and  when  such  has  been  provided, 
the  crank-pin  will  also  be  strong  enough  for  the  work  it  has  to  do.  These  remarks 
again  apply  to  the  main-rod  journal  of  the  crank-pins,  in  which  this  journal  is  the 
outer  one,  as  shown  in  Fig.  487 ;  they  also  apply  to  the  side-rod  crank-pin,  such 
as  is  shown  in  Fig.  486 ;  but  they  do  not  apply  to  side-rod  journal  on  the  main  crank- 
pins,  as  will  presently  be  seen. 

Since  the  pressure  per  square  inch  of  the  projected  area  of  a  steel  crank-pin  journal 
should  not  exceed  1,600  pounds,  we  readily  find,  by  Eule  74,  the  projected  area  of  the 
main-rod  journal  for  consolidation,  Mogul,  and  ten-wheeled  engines.  Thus : 


MODERN  LOCOMOTIVE   CONSTRUCTION.  399 

EXAMPLE  110. — What  should  be  the  projected  area  of  the  main-rod  journal  on  the 
main  crunk-pin  for  a  Mogul  engine  having  cylinders  18  inches  diameter?  Maximum 
steam  pressure  on  the  piston  is  130  pounds  per  square  inch. 

Multiplying  the  area  in  square  inches  of  the  piston  by  the  steam  pressure  per 
square  inch,  we  have 

254.47  x  130  =  33081.10  pounds, 

which  is  the  maximum  steam  pressure  on  the  piston. 
According  to  Rule  74,  we  have 

33081.10 

.  .  =  20.67  square  inches, 

which  is  the  projected  area  of  the  main-rod  crank-pin  journal. 

Before  we  can  find  the  diameter  and  the  length  of  this  journal  we  must  establish 
a  ratio  between  the  diameter  and  the  length.  There  are  a  few  engines  of  the  classes 
now  under  consideration  in  which  the  diameter  of  the  main-rod  crank-pin  journal 
exceeds  its  length ;  and  on  the  other  hand  we  meet  with  a  few  engines  in  which  the 
diameter  is  less  than  the  length.  But  in  a  large  majority  of  these  engines  the  diame- 
ter of  the  main-rod  crank-pin  journal  is  equal  to  its  length.  We  shall  therefore  adopt 
a  rule  by  which  we  can  find  the  diameter  and  the  length  which  are  equal  to  each 
other.  To  the  diameter  so  found  we  shall  add,  for  wear,  about  iV  or  £  inch,  so 
as  to  obtain  such  diameters  whose  fractions,  if  they  have  any,  are  divisible  by  £  inch ; 
as  for  the  length,  we  shall  adopt  the  nearest  which  contains  fractions  that  can  be 
divided  by  J  inch,  and  this  may  be  either  a  little  less  or  a  little  greater  than  the 
length  found  by  the  following  rule : 

RULE  78. — To  find  the  diameter  and  length  of  the  main-rod  crank-pin  journal : 
Extract  the  square  root  of  the  projected  area,  as  found  by  Rule  74.  If  this  square  root 
does  not  contain  any  fractions,  or  if  it  does  contain  fractions  which  can  be  divided  by 
i  inch,  add  to  it  i  of  an  inch  for  wear ;  the  sum  will  be  the  diameter.  If  the  square 
root  contains  fractions  which  cannot  be  divided  by  J  inch,  add  ^  or  a  little  more,  so  as 
to  obtain  diameters  which  are  divisible  by  £  inch.  The  square  root  of  the  projected 
area  will  also  be  the  length  of  the  pin ;  if  it  is  not  divisible  by  £  inch,  adopt  the  near- 
est length  which  is  divisible  by  &  inch. 

EXAMPLE  111. — Find  the  diameter  and  length  of  the  main-rod  crank-pin  journal 
for  a  Mogul  engine  having  cylinders  18  inches  diameter,  maximum  steam  pressure  on 
piston  130  pounds  per  square  inch. 

In  Example  110  we  have  found  that  the  projected  area  of  this  journal  is  20.67 
square  inches.  The  square  root  of  20.67  is  4.54  inches,  which  is  a  little  more  than 
4£  inches,  hence  the  diameter  is  4g  inches ;  and  the  length  is  4J  inches. 

344.  The  projected  area  of  the  side-rod  journal  of  the  main  crank-pin,  that  is,  the 
inner  journal  in  Fig.  487,  as  made  by  different  builders,  varies  somewhat ;  hence  we 
find  that  in  a  number  of  engines  the  projected  area  of  the  side-rod  journal  is  greater 
than  that  of  the  main-rod  journal ;  and  in  a  number  of  engines  it  is  less;  but  in  the 
majority  of  engines  the  projected  areas  of  the  two  journals  are  equal,  and  these  lalter 
proportions  we  shall  adopt.  Hence  the  projected  area  of  the  side-rod  journal  of  the 
main  crank-pin  is  found  by  Rule  74. 


330  MODERN  LOCOMOTIVE   CONSTRUCTION. 

True,  if  we  had  only  to  make  provisions  for  the  prevention  of  heating,  a  smaller 
area  would  suffice,  but,  since  the  outer  journal  of  this  pin  is  subjected  to  a  greater 
pressure  than  is  brought  to  bear  on  the  side-rod  journal,  and  since  the  pressure  on  the 
outer  pin  acts  with  a  leverage,  we  require  a  large  projected  area  so  as  to  provide  for 
the  strength  of  the  pin. 

The  ratio  between  the  diameter  and  length  of  this  side-rod  journal,  as  made  by 
different  builders,  also  varies ;  but  the  most  common  practice  is  to  make  the  diameter 
1J  times  larger  than  the  length;  and  these  proportions  we  shall  adopt.  We  have 
therefore  the  following  rule : 

RULE  79. — To  find  the  length  and  diameter  of  the  side-rod  journal  for  the  main 
crank-pin,  when  this  journal  is  next  to  the  wheels,  divide  the  projected  area,  as  found 
by  Rule  74,  by  5 ;  multiply  the  quotient  by  4,  and  extract  the  square  root  from  the 
product ;  this  square  root  will  be  the  length  of  the  side-rod  journal.  If  this  length 
is  not  divisible  by  £  inch,  adopt  the  nearest  one  which  can  be  divided  by  £  inch. 
For  the  diameter  multiply  the  square  root  found  as  above  by  1.25 ;  if  this  product  is 
divisible  by  ^  inch,  add  to  it  |  inch  for  wear ;  the  sum  will  be  the  required  diameter ; 
if  this  product  is  not  divisible  by  J-,  add  from  tV  to  £  inch,  so  as  to  make  it  divisible 
by  £  inch ;  the  sum  will  be  the  required  diameter. 

EXAMPLE  112. — Find  the  length  and  diameter  of  the  side-rod  journal  for  the  main 
crank-pin  for  a  Mogul  engine  having  cylinders  18  inches  diameter ;  maximum  steam 
pressure  on  the  piston,  130  pounds  per  square  inch.  The  projected  area  of  this  journal 
is,  according  to  Rule  74,  20.67  square  inches ;  in  fact,  this  area  we  have  found  in 
Example  110. 

According  to  Rule  79,  we  have 

"*>  *  *  =  16.53. 


5 
The  square  root  of  16.53  is 

V16.53  =  4.06  inches. 

Hence  the  length  of  this  journal  is  4  inches. 

Again,  for  the  diameter  we  have  4.06  x  1.25  =  5.07  inches,  which  is  a  little  more 
than  5  inches ;  hence  the  diameter  of  this  journal  is  5£  inches. 

In  a  similar  way  we  can  determine  the  dimensions  of  the  journals  for  any  main 
crank-pin  (made  of  steel)  under  different  pressures  for  ten-wheeled,  Mogul,  and  con- 
solidation engines. 

The  dimensions  of  the  main  crank-pin  journals  given  in  the  following  tables  have 
been  computed  by  the  foregoing  rules. 


MODERN  LOCOMOTIVE  CONSTRUCTION. 

TABLE  42. 


331 


DIMENSIONS   OF   THE  MAIN   CRANK-PIN  JOURNALS  (STEEL)  FOR  MOGUL,  TEN-WHEELED,  AND   CONSOLIDA- 
TION  ENGINES.      MAXIMUM   STEAM  PRESSURE  ON  THE   PISTON,   120   POUNDS   PER   SQUARE  INCH. 


Diameter  of  Cylin- 
ders. 

Main-rod  Journals. 

Side-rod  Journals. 

Diameter. 

Length. 

Diameter. 

Length. 

9 

24  inches. 

24  inches. 

24  inches. 

2    inches. 

10 

24 

2f 

2t 

24 

11 

21 

y 

3 

2| 

12 

3 

21 

3f 

§ 

13 

34 

34 

3f 

21 

14 

34 

3f 

31 

3 

15 

3t 

3» 

44 

31 

16 

4 

31 

44 

34 

17 

4i 

4* 

4* 

3f 

18 

44 

4f 

5 

31 

19 

4} 

4* 

54 

44 

20 

5 

4* 

54 

4| 

21 

54 

54 

H 

44 

22 

5* 

5f 

6 

4* 

TABLE  43. 

DIMENSIONS  OP  THE  MAIN  CRANK-PIN  JOURNALS  (STEEL)  FOR  MOGUL,  TEN-WHEELED,  AND  CONSOLIDA- 
TION  ENGINES.      MAXIMUM  STEAM  PRESSURE   ON  THE  PISTON,    130   POUNDS  PER  SQUARE   INCH. 


Diameter  of  (rylin- 
den. 

Main-rod  Journals. 

Side-rod  Jonrnale. 

Diameter. 

Length. 

Diameter. 

Length. 

9 

2{  inches. 

2J  inches. 

2|  inches. 

2    inches. 

10 

§ 

24 

21 

24 

11 

21 

21 

34 

24 

12 

34 

3 

34 

21 

13 

3f 

34 

31 

3 

14 

3i 

34 

4 

34 

15 

31 

3} 

44 

3f 

16 

44 

4 

4* 

3* 

17 

4* 

44 

*1 

31 

18 

«i 

44 

54 

4 

19 

41 

41 

54 

44 

20 

54 

5 

5f 

44 

21 

Sf 

54 

6 

4} 

22 

5» 

54 

64 

5 

TABLE   44. 

DIMENSIONS  OF  THE  MAIN  CRANK-PIN  JOURNALS  (STEEL)  FOR  MOGUL,  TEN-WHEELED,  AND  CONSOLIDA- 
TION  ENGINES.      MAXIMUM   STEAM   PRESSURE  ON   THE   PISTON,    140   POUNDS   PER  SQUARE   INCH. 


Diameter  of  Cylin- 
ders. 

Main-rod  Journals. 

Side-rod  Journal*. 

Diameter. 

Length. 

Diameter. 

Length. 

9 

24  inches. 

2}  inches. 

21  inches. 

24  inches. 

10 

2} 

1 

2* 

3 

§ 

11 

3 

< 

21 

34 

24 

12 

34 

34 

3| 

2} 

13 

34 

3* 

31 

3 

14 

31 

3| 

44 

34 

15 

4 

31 

44 

34 

16 

44 

44 

41 

31 

17 

44 

4» 

5 

4 

18 

41 

4| 

5f 

44 

19 

r>4 

5 

5* 

44 

20 

5f 

54 

51 

4t 

21 

5t 

54 

64 

41 

22 

51 

51 

64 

54 

332 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


TABLE   45. 

DIMENSIONS  OF   THE   MAIN   CRANK-PIN  JOURNALS  (STEEL)  FOR  MOGUL,  TEN-WHEELED,  AND  CONSOLIDA- 
TION ENGINES.      MAXIMUM   STEAM   PRESSURE   ON   THE   PISTON,   150   POUNDS   PER  SQUARE  INCH. 


Diameter  of  Cylin- 
ders. 

Main-rod  Journals. 

Side-rod  Journals. 

Diameter. 

Length. 

Diameter. 

Length. 

9 

24  inches. 

2f  inijhes. 

2f  inches. 

24  inches. 

10 

H 

It 

34 

24 

11 

34 

3 

3| 

24 

12 

3f 

3* 

3J 

24 

13 

34 

34 

4 

34 

14 

3J 

34 

*i 

3f 

15 

44 

4 

*4 

34 

16 

« 

4f 

41 

3* 

17 

4| 

4* 

BJ 

44 

18 

5 

4* 

51 

4f 

19 

5* 

8* 

Bf 

44 

20 

54 

5f 

64 

H 

21 

Bf 

54 

6| 

54 

22 

64 

6 

64 

5f 

TABLE  46. 

DIMENSIONS   OF   THE   MAIN   CRANK-PIN   JOURNALS  (STEEL)  FOR  MOGUL,  TEN-WHEELED,  AND  CONSOLIDA- 
TION  ENGINES.      MAXIMUM   STEAM  PRESSURE   ON  THE   PISTON,   160   POUNDS   PER   SQUARE  INCH. 


Diameter  of  Cylin- 
ders. 

Main-rod  Journals. 

Side-rod  Journals. 

Diameter. 

Length. 

Diameter. 

Length. 

9 

24  inches. 

24  inches. 

2|  inches. 

2J  inches. 

10 

2J 

2f 

3± 

24 

11 

34 

3 

34 

2f 

12 

31 

3f 

3J 

3 

13 

3f 

34 

44 

3i 

14 

4 

34 

44 

34 

15 

4* 

44 

4J 

3i 

16 

*i 

4| 

54 

4 

17 

*i 

4* 

54 

4t 

18 

54 

5 

5t 

44 

19 

BJ 

5f 

6 

4i 

20 

51 

54 

6| 

5 

21 

6 

51 

64 

54 

22 

61 

64 

7 

54 

There  are  very  few  Mogul  engines  built  with  cylinders  9  inches  diameter,  and  we 
do  not  know  of  any  consolidation  engines  with  cylinders  of  such  small  diameter;  it 
may  therefore  appear  unnecessary  to  extend  the  tables  to  such  small  cylinders.  But 
we  have  seen  that,  in  narrow-gauge  passenger  engines,  the  side-rods  are  often  placed 
next  to  the  driving  wheels ;  these  engines  generally  have  small  cylinders,  consequently 
these  tables  are  useful  for  obtaining  the  dimensions  of  main  crank-pins  for  this  class, 
and  other  classes  of  engines  whose  side-rods  are  placed  next  to  the  driving  wheels. 


SIDE-ROD   PINS   FOE   TEN-WHEELED   AND   MOGUL  ENGINES. 

345.  In  comparing  the  side-rod  pins  for  ten-wheeled  and  Mogul  engines  with  the 
side-rod  pins  for  eight-wheeled  passenger  engines  whose  cylinders  are  equal  in  size  to 
those  in  the  former  classes  of  engines,  and  all  subjected  to  the  same  steam  pressure, 


MODERN   LOCOMOTirE    CONSTRUCTION.  333 

we  find  that  the  side-rod  pins  in  ten-wheeled  and  Mogul  engines  are  smaller  than  those 
in  eight-wheeled  passenger  engines.  The  reason  for  this  is  that  in  passenger  engines 
we  have  only  two  driving  wheels  on  each  side,  and  in  ten-wheeled  and  Mogul  engines 
we  have  three  driving  wheels  on  each  side.  With  cylinders  of  equal  size,  and  equal 
steam  pressures,  the  thrust  on  the  main-rod  crank-pin  journals  in  all  the  different 
classes  of  engines  will  be  practically  equal.  Now  assuming  that  the  total  weight  on 
the  driving  wheels  is  equally  distributed,  the  pressure  on  the  side-rod  pins  in  ten- 
wheeled  and  Mogul  engines  must  be  less  than  that  on  the  side-rod  pins  in  passenger 
engines,  because  in  the  latter  class  nearly  all  the  pressure  on  the  main  pin  is 
transmitted  to  two  wheels,  whereas,  in  the  former  class  an  equal  amount  of  pressure 
is  transmitted  to  three  wheels.  Hence,  the  rules  previously  given  for  determining  the 
dimensions  of  the  side-rod  pin  for  eight-wheeled  passenger  engines  are  not  suitable  for 
finding  the  dimensions  of  the  side-rod  pins  for  ten-wheeled  and  Mogul  engines. 

The  following  rules  will  give  results  which  agree  closely  with  the  average  good 
practice : 

KULE  80. — Divide  the  total  steam  pressure  on  the  piston  by  2,800 ;  the  quotient 
will  be  the  number  of  square  inches  in  the  projected  area  of  steel  side-rod  pins  in  the 
front  and  rear  driving  wheels  under  ten-wheeled  and  Mogul  engines. 

RULE  81. — To  find  the  diameter  and  length  of  the  front  and  rear  side-rod  pins  in 
the  above  classes  of  engines,  extract  the  square  root  of  the  projected  ai'ea,  as  found  by 
Rule  80.  If  this  square  root  does  not  contain  any  fractions,  or  if  it  does  contain 
fractions  which  can  be  divided  by  £  inch,  add  to  it  £  inch  for  wear ;  the  sum  will 
be  the  diameter  of  the  pins.  If  the  square  root  contains  fractions  which  cannot  bo 
divided  by  £  inch,  add  about  -fa  inch,  or  a  little  more,  so  as  to  obtain  diameters  which 
are  divisible  by  J  inch.  The  square  root  of  the  projected  area  will  also  be  the  length 
of  the  pins;  if  it  is  not  divisible  by  £  inch,  adopt  the  nearest  length  which  is  divisible 
by  i  inch. 

EXAMPLE  113. — Find  the  diameter  and  length  of  the  front  and  rear  side-rod  pins  in 
a  Mogul  engine  having  cylinders  18  inches  diameter ;  maximum  steam  pressure  on  the 
piston,  130  pounds  per  square  inch. 

The  total  steam  pressure  on  the  piston  is  equal  to  its  area  multiplied  by  the  steam 
pressure  per  square  inch ;  therefore, 

254.47  x  130  =  33081.10  pounds  =  total  pressure  on  the  piston. 
According  to  Rule  80,  the  projected  area  is  equal  to 

33081.10 

— 9800       :  11.81  square  inches. 

According  to  Rule  81,  we  must  find  the  square  root  of  11.81,  which  is  3.43,  or,  we 
may  say  it  is  practically  equal  to  3^$  inches;  hence,  the  diameter  of  this  pin  is  3£ 
inches  and  its  length  3fj  inches. 

In  a  similar  manner  the  dimensions  of  steel  side-rod  pins  for  ten-wheeled  and 
Mogul  engines  in  the  following  tables  have  been  computed : 


334 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


TABLE  47. 

DIMENSIONS    OF    STEEL    SIDE-BOD    PINS    IN    THE   FRONT    AND    REAR  WHEELS    FOR  TEN- WHEELED  AND 
MOGUL  ENGINES.     .MAXIMUM   STEAM   PRESSURE   ON  THE  PISTON,    120   POUNDS   PER  SQUARE  INCH. 


Diameter  of  Cylinders. 

Diameter  of  Journals. 

Length  of  Journals. 

9 

14  inches. 

1}  inches. 

10 

2          ' 

14       " 

11 

2*        ' 

2         " 

12 

2|        ' 

2*       " 

13 

2*        ' 

2|       " 

14 

2f        ' 

21       " 

15 

24        ' 

2f       " 

16 

34        ' 

3 

17 

3i        ' 

34       " 

18 

34       " 

3|       " 

19 

3$       " 

3|       " 

20 

3f       " 

3f       " 

21 

4         " 

34       " 

22 

44       " 

4         " 

TABLE  48. 

DIMENSIONS    OF    STEEL    SIDE-ROD    PINS    IN    THE    FRONT    AND    REAR    WHEELS    FOR   TEN-WHEELED   AND 
MOGUL  ENGINES.      MAXIMUM  STEAM   PRESSURE   ON  THE  PISTON,   130   POUNDS  PER  SQUARE  INCH. 


Diameter  of  Cylinders. 

Diameter  of  Journals. 

Length  of  Journals. 

9 

14  inches. 

If  inches 

10 

2 

14       " 

11 

2±       " 

24      " 

12 

2|        | 

2J       " 

01                It 

13 

*^8 

^4 

14 

2f        ' 

2f      " 

15 

3          ' 

24       ' 

16 

34        ' 

3          ' 

17 

3f        ' 

3±        ' 

18 

34        ' 

3f        ' 

19 

3}        ' 

3*        ' 

20 

34 

3f        ' 

21 

44    •; 

4 

22 

4± 

4i 

TABLE  49. 

r  * 

DIMENSIONS    OF    STEEL   SIDE-ROD    PINS    IN    THE    FRONT    AND   REAR    WHEELS    FOR   TEN-WHEELED   AND 
MOGUL  ENGINES.      MAXIMUM   STEAM   PRESSURE   ON  THE   PISTON,    140   POUNDS   PER  SQUARE  INCH. 


Diameter  of  Cylinders. 

Diameter  of  Journals. 

Length  of  Journals. 

9 

14  inches. 

If  inches. 

10 

24       " 

2         " 

11 

2i       " 

24        ' 

12 

1>4         " 

21        ' 

13 

'1$        " 

24         ' 

14 

24        " 

2f        ' 

15 

34        " 

3          ' 

16 

3i       " 

34         ' 

17 

34       " 

3|        ' 

18 

3f       " 

34        ' 

19 

34       " 

3f        ' 

20 

44       " 

4 

21 

4i       " 

44       " 

22 

4|       " 

4|       " 

MODERN  LOCOMOT/rt;  COXSTRVCTIOX. 


335 


TABLE  50. 

DDffiNSIONS    OP    STEEL    SIDE-BOD    PINS    IN    THE    FRONT    AND    REAR   WHEELS    FOR    TEN-WHEELED    AND 
MOGUL  ENGINES.      MAXIMUM  STEAM  PRESSURE  ON  THE  PISTON,   150   POUNDS  PER  SQUARE  INCH. 


Diameter  of  Cylinders. 

Diameter  of  Journals. 

Length  of  Journals. 

9 

2    inches. 

H  inches. 

10 

2i       " 

2        » 

11 

24        ' 

2J      " 

12 

2*        ' 

2i      '• 

13 

2t       ' 

24      " 

14 

3         ' 

2*      " 

15 

3*        ' 

3        " 

16 

3f        ' 

3*       ' 

17 

34        ' 

3i       • 

18 

3f      " 

34        ' 

19 

4        " 

3J       ' 

20 

4i      " 

44       ' 

21 

4i      " 

4|        • 

22 

44      " 

44       ' 

TABLE  51. 

DDIENSIONS    OF    STEEL    SIDE-ROD    PINS   IN    THE    FRONT    AND    REAR    WHEELS    FOR    TEN-WHEELED   AND 
MOGUL    ENGINES.      MAXIMUM  STEAM  PRESSURE  ON  THE  PISTON,  160  POUNDS  PER  SQUARE  INCH. 


Diameter  of  Cylinders. 

Diameter  of  Journals. 

Length  of  Journals. 

9 

2^  inches. 

2    inches. 

10 

2i       " 

24      " 

11 

24       " 

2|      " 

12 

24       " 

24      " 

13 

2*       " 

2f        ' 

14 

34      " 

3          ' 

15 

3±      " 

34        ' 

16 

34      " 

31        ' 

17 

3|      " 

3+        ' 

18 

3J      " 

3J        ' 

19 

4*       " 

4         " 

20 

4!       « 

4i       " 

21 

41       " 

44      " 

22 

4f       " 

44      " 

SIDE-BOD  PINS  FOR  CONSOLIDATION   ENGINES. 

346.  In  the  foregoing  article  we  have  seen  that  the  side-rod  pins  for  eight-wheeled 
passenger  engines  are  larger  than  those  for  Mogul  and  ten-wheeled  engines.  If,  now, 
we  compare  the  relative  pressures  on  the  side-rod  pins  in  Mogul  and  consolidation 
engines,  we  will  find  that  for  the  latter  class  of  engines  we  may  make  the  side-rod 
pins  still  smaller ;  that  is,  say,  they  may  be  made  smaller  than  those  in  Mogul  engines. 
The  reason  for  this  will  be  seen  by  referring  to  Fig.  488,  which  shows  the  arrange- 
ments of  the  driving  wheels  on  one  side  of  a  consolidation  engine.  The  main-rod  pin 
is  marked  M,  and  A,  B,  D  are  the  side-rod  pins.  For  the  purpose  of  comparison,  we 
may  assume  that  in  all  locomotives  the  total  weight  on  the  drivers  is  equally  distributed 
on  the  wheels;  and  we  may  also  assume  that  the  whole  pressure  on  the  main-rod 
crank-pin  journal  is  required  for  turning  the  wheels  (this,  of  course,  is  not  exactly  true). 
Under  these  conditions  it  will  require  one-fourth  of  the  pressure  on  the  main  crank-pin 


336  MODERN  LOCOMOTIVE   CONSTRUCTION. 

journal  to  turn  each  wheel.  In  Mogul  engines  we  have  only  three  driving  wheels 
on  each  side  of  the  engine ;  and  with  cylinders  of  the  same  size  as  those  in  a  consoli- 
dation engine,  and  also  equal  steam  pressures,  the  thrust  on  the  main-rod  crank-pin 
journal  will  practically  be  the  same  in  both  engines ;  and  the  pressure  on  the  side-rod 


Fourth  Vair  of  Drivers          Third  Pair  of  Drivers          Second  Fatr  of  Drivers  First  Pair  of  Drivers 


Centrt  of  Slain  Hod 
M 

Fig.  488 

pins  for  Mogul  engines  will  be  equal  to  one-third  of  that  on  main-rod  journal  instead 
of  one-fourth,  as  in  consolidation  engines.  Therefore,  since  the  pressure  on  the  side- 
rod  pins  in  the  latter  class  of  engines  is  less  than  that  in  the  former  classes,  it  follows 
that  the  side-rod  pins  for  consolidation  engines  may  also  be  reduced  in  size.  True, 
these  pins  are  subjected  to  other  pressures  besides  that  due  to  pressure  on  the  piston  ; 
but  these  may  be  provided  for  by  choosing  a  proper  divisor  for  the  total  steam 
pressure  on  the  piston  in  determining  the  projected  area,  as  we  have  done  in  the 
following  rule  : 

RULE  82.  —  Divide  the  total  maximum  steam  pressure  on  the  piston  by  3,200;  the 
quotient  will  be  the  number  of  square  inches  in  the  projected  area  of  any  one  of  the 
steel  side-rod  pins  A,  B,  D  (Fig.  488)  for  consolidation  engines. 

The  length  and  diameter  of  these  pins  are  equal,  or  nearly  so,  in  the  majority  of 
engines  of  this  class  ;  hence  the  following  rule  : 

RULE  83.  —  To  find  the  length  and  diameter  of  steel  side-rod  pins  for  consolidation 
engines,  extract  the  square  root  of  the  projected  area  as  found  by  Rule  82  ;  result  will 
be  the  length  of  the  pin  ;  if  this  length  contains  a  fraction  not  divisible  by  £  inch, 
then  adopt  the  nearest  length  which  can  be  divided  by  £  inch.  For  the  diameter,  add 
to  the  length  found  £  inch  for  wear  ;  the  sum  will  be  the  required  diameter. 

EXAMPLE  114.  —  Find  the  dimensions  of  the  side-rod  pins  for  a  consolidation 
engine  having  cylinders  20  inches  diameter  ;  maximum  steam  pressure  on  the  piston, 
130  pounds  per  square  inch. 

The  area  of  a  piston  20  inches  diameter  is  314.16  square  inches;  hence  the 
maximum  steam  pressure  on  the  piston  will  be  314.16  x  130  =  40840.8  pounds. 

40840  8 
According  to  Rule  82,  we  have     Qonf  -  =  12.76  square  inches,  which  is  the  projected 


area  of  the  pin.  The  square  root  of  12.76  is  3.57.  Hence,  the  length  of  the  side-rod 
pin  will  be  3£  inches,  and  its  diameter  will  be  3§  inches. 

The  dimensions  given  in  the  following  tables  have  been  computed  by  the  fore- 
going rules. 

There  is  a  possibility,  when  a  consolidation  engine  is  running  over  an  uneven 
track,  causing  the  wheels  to  run  out  of  alignment,  that  the  side-rod  pin  B  will  be 
subjected  to  a  greater  pressure  than  the  side-rod  pins  A  and  D.  Consequently,  we 
frequently  find  the  side-rod  pin  B  made  from  i  to  §  of  an  inch  greater  in  diameter  than 
that  of  the  pins  A  and  D.  This  is  good  practice,  but  not  a  universal  one,  as  in  many 


LOCOMOTIVE  CONSTRUCTION. 


337 


engines  all  the  side-rod  pins  are  the  same  size.  In  computing  the  dimensions  given  in 
these  tables,  we  have  assumed  that  all  the  side-rod  pins  should  be  equal  in  size.  If 
it  is  desirable  to  increase  the  size  of  the  pin  B,  its  diameter  only  should  be 
increased ;  the  length  should  remain  as  given ;  by  so  doing  the  center  lines  of  the  side- 
rods  can  be  kept  more  readily  in  the  same  vertical  plane,  which  is  of  importance. 


TABLE   52. 

DIMENSIONS  OF   STEEL   SIDE-ROD  PINS  FOR  CONSOLIDATION   ENGINES.      MAXIMUM  STEAM   PRESSURE 
ON  THE  PISTON,    120   POUNDS   PER   SQUARE   INCH. 


Diameter  of  Cylinder*. 

Diameter  of  Journals. 

Length  of  Journals. 

14 

24  inches. 

2}  inches. 

15 

24       " 

24       " 

16 

24      " 

2}       " 

17 

3         " 

24       " 

18 

34      " 

3         " 

19 

3|       " 

3J       " 

20 

3f       " 

34       " 

21 

31       " 

3f       " 

22 

34       " 

3f       " 

TABLE    53. 

DIMENSIONS  OF  STEEL  SIDE-ROD  PINS   FOR  CONSOLIDATION  ENGINES.      MAXIMUM  STEAM  PRESSURE 
ON  THE  PISTON,    130   POUNDS   PER  SQUARE   INCH. 


Diameter  of  Cylinders. 

Diameter  of  Journals. 

Length  of  Journals. 

14 

2f  inches. 

24  inches. 

15 

2J       " 

2f       " 

16 

3         " 

24       " 

17 

34       " 

3         " 

18 

3f        ' 

34       " 

19 

34        ' 

3f       " 

20 

3f        ' 

31       " 

21 

34        ' 

3f       " 

22 

44      ' 

4         " 

TABLE    54. 

DDIENSIONS   OK   STEEL  SIDE-ROD   PINS   FOR  CONSOLIDATION  ENGINES.      MAXIMUM  STEAM  PRESSURE 
ON  THE   PISTON,    140   POUNDS   PER  SQUARE   INCH. 


Diameter  of  Cylinders. 

Diameter  of  Journals. 

Length  of  Journals. 

14 

-:•  inches. 

24  inches. 

15 

24       " 

2f      " 

16 

3*        ' 

3        " 

17 

34        ' 

34       " 

18 

34        ' 

3|       " 

19 

31        ' 

34      " 

20 

3t        ' 

3*       " 

21 

4 

34       " 

22 

44       " 

44      " 

338 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


TABLE    55. 

DIMENSIONS   OF   STEEL   SIDE-ROD  PINS   FOR  CONSOLIDATION   ENGINES.      MAXIMUM   STEAM  PRESSURE 
ON  THE   PISTON,   150   POUNDS  PER   SQUARE   INCH. 


Diameter  of  Cylinders. 

Diameter  of  Journals. 

Length  of  Journals. 

14 

2f  inches. 

2|  inches. 

15 

3 

2* 

16 

34 

3 

17 

3| 

3i 

18 

3f 

8* 

19 

3f 

3| 

20 

4 

3J 

21 

44 

4 

22 

4f 

*J 

TABLE   56. 

DIMENSIONS   OF   STEEL  SIDE-ROD   PINS  FOR  CONSOLIDATION   ENGINES.      MAXIMUM  •  STEAM   PRESSURE 
ON   THE  PISTON,    160   POUNDS  PER  SQUARE  INCH. 


Diameter  of  Cylinders. 

Diameter  of  Journals. 

Length  of  journals. 

14 

2|  inches. 

2J  inches. 

15 

34 

3         ' 

16 

3i 

34       ' 

17 

34 

3| 

18 

3* 

34 

19 

3J 

8| 

20 

44 

4 

21 

4i 

44 

22 

44 

4f 

WHEEL  FITS. 

We  have  seen  that  in  eight-wheeled  passenger  engines  (two  driving  wheels  on 
each  side)  the  diameter  of  the  wheel  fit  of  the  side-rod  pin  is  equal  to  that  of  the  main 
pin.  In  ten-wheeled,  Mogul,  and  consolidation  engines,  the  diameter  of  the  wheel  fit 
of  the  side-rod  pins  is  less  than  that  of  the  main  pin ;  it  is  generally  from  i  to  £  inch 
greater  than  the  diameter  of  the  journal.  The  wheel  fit  for  the  main  pin  should  never 
be  less  than  the  diameter  of  the  journal  next  to  the  wheels ;  in  fact,  it  is  better  practice 
to  increase  it  from  £  to  J  inch. 


KNUCKLE  JOINTS. 


347.  The  side-rods  in  all  locomotives  which  have  more  than  two  driving  wheels  on 
each  side  have  a  knuckle  joint ;  its  form  in  many  engines  is  similar  to  that  shown  in 
Figs.  489  and  490.  In  Mogul  and  ten-wheeled  engines  there  is  only  one  knuckle  joint  on 
each  side  of  the  engine ;  sometimes  it  is  placed  in  the  rear  of  the  main  pin,  as  shown 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


in  Fig.  437  (Art.  309),  and  occasionally  we  find  it  placed  in  front  of  the  main  pin ; 
there  is  an  advantage  in  placing  it  in  the  rear  (see  Art.  309).  In  all  consolidation 
engines  one  knuckle  joint  is  placed  in  the  rear  of  the  main  pin  and  another  one  in 
front  of  the  side-rod  pin  B  (Fig.  488),  or  in  general  they  are  placed  in  the  rear  of  and 


--—-4- 


i 


i 


JUg.  400 


I 


close  to  the  pins  in  the  third  pair  of  drivers ;  and  in  the  front  of  and  close  to  the 
pins  in  the  second  pair  of  drivers.  The  solid  ends  b  I  (Fig.  490)  of  the  knuckle 
joints  are  generally  forged  to  the  central  side-rod  straps,  and  the  forked  ends  are 
forged  to  the  front  and  rear  side-rods. 

The  center  lines  of  the  side-rods  always  lie  in  one  vertical  plane,  represented  by 
the  line  C  d  in  Fig.  488.  This  plane  is  also  represented  by  the  line  c  d  in  Fig.  490.  In 
this  figure  it  will  be  noticed  that  the  center  line  b  b  of  the  knuckle  joint  does  not  coin- 
cide with  the  center  line  c  d  of  the  side-rods;  the  reason  for  this  is,  that  in  a  great 
number  of  engines  the  inner  face /of  the  flange  of  the  main-rod  brass  is  close  to  the 


340 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


outer  face  of  the  hub  of  wheel,  consequently  the  inner  face  e  of  the  knuckle  joint 
must  not  project  beyond  the  face/  and  therefore  the  center  b  b  of  the  forked  end  will 
often  be  out  of  line  with  center  c  d  of  the  rod. 

348.  The  form  of  the  pin  through  the  knuckle  joint  is  plainly  shown  in  Fig.  490. 
A  larger  portion  of  its  head  h  is  made  conical,  and  the  remainder,  g,  cylindrical ;  the 
diameter  of  this  cylindrical  portion  is  generally  a  little  greater  than  the  diameter  of 
the  body  of  the  pin,  so  that  the  latter  can  be  passed  easily  through  the  hole  bored 
for  the  head.  Another  form  of  the  knuckle-joint  pin  is  shown  in  Fig.  491 ;  here  the 
holes  in  both  wings  a  a2  of  the  forked  end  are  reamed  out  with  one  tapered  reamer ; 
this  plan  seems  to  be  the  most  popular  one. 

The  ratio  between  the  length  and  the  diameter  of  the  pin,  as  made  by  different 
builders,  varies  somewhat,  as  will  be  seen  by  referring  to  Figs.  490  and  491 ;  but  we 
believe  good  results  will  be  obtained  by  making  the  diameter  25  per  cent,  greater  than 
the  length  of  the  pin. 

Generally,  the  pins  in  the  knuckle  joints  are  made  of  wrought-iron,  and  are  case 
hardened.  Iron  is  to  be  preferred  to  steel,  as  iron  pins,  particularly  when  their  form 
is  like  that  shown  in  Fig.  491,  are  not  as  liable  to  break  off  in  the  shank  as  steel  ones. 
In  many  engines  these  pins  work  in  wrought-iron  bushings  case  hardened ;  in  a  few 
instances  they  work  in  brass  bearings,  arranged  as  shown  in  Figs.  441,  445. 

The  knuckle-joint  pins  should  be  prevented  from  turning  in  their  outer  bearings ; 
for  this  purpose  dowel  pins  are  inserted,  as  shown  at  p,  Fig.  491. 

So  long  as  the  side-rods  remain  in  perfect  alignment,  the  knuckle-joint  pins  do 
not  rotate  in  their  middle  bearings ;  under  other  conditions  the  amount  of  rotation  is 


Fig.  491 


\e TJ-dia. 1 


L 0H 

"•IT 


very  small.  Hence  their  liability  of  heating  is  not  so  great  as  that  of  the  crank-pins ; 
consequently,  the  pressure  per  square  inch  of  projected  area  of  a  knuckle-joint  pin  can 
be  considerably  greater  than  the  pressure  per  square  inch  of  projected  area  of  a  crank- 
pin.  Careful  observation  seems  to  indicate  that  the  projected  area  of  a  knuckle-joint 
pin  may  be  at  once  determined  from  the  pressure  on  the  piston,  and  that  for  every 
7,000  pounds  of  the  maximum  steam  pressure  on  the  piston  one  square  inch  of 
projected  area  should  be  allowed.  Hence,  for  obtaining  the  length  and  diameter  of  a 
knuckle-joint  pin,  we  have  the  following  rules : 

RULE  84. — Divide  the  total  maximum  steam  pressure  on  the  piston  by  7,000 ;  the 
quotient  will  be  the  number  of  square  inches  in  the  projected  area  of  the  pin. 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


341 


RULE  85. — To  find  the  length  and  diameter  of  the  knuckle-joint  pin,  divide  the 
projected  area,  as  found  by  Eule  84,  by  5,  multiply  the  quotient  by  4,  and  extract  the 
square  root  of  the  product;  the  result  will  be  the  length  of  the  pin.  Multiply  the 
length  by  1.25 ;  the  product  will  be  the  diameter  of  the  pin.  Only  fractions  divisible 
by  -fa  inch  are  adopted ;  hence,  if  the  dimensions  found  by  calculation  contain  frac- 
tious not  divisible  by  -fa  inch,  adopt  the  nearest  one  which  can  be  so  divided. 

EXAMPLE  115. — It  is  required  to  find  the  dimensions  of  a  knuckle-joint  pin  for  a 
consolidation  engine  having  cylinders  20  inches  diameter ;  maximum  steam  pressure 
on  the  piston,  140  pounds  per  square  inch. 

The  maximum  pressure  on  the  piston  will  be 

314.16  x  140  =  43982.4  pounds. 
According  to  Rule  84,  we  have 

43982.4 

=  6.283+  square  inches  in  the  projected  area. 


According  to  Rule  85,  we  have 

6.283 


x  4  =  5.02+. 


The  square  root  of  5.02  is  V5.02  =  2.24  inches,  which  is  the  length  of  the  pin.  And 
the  diameter  will  be 

2.24  x  1.25  =  2.80  inches. 

Adopting  the  nearest  fractions  which  are  divisible  by  ^  inch,  we  have  for  the  length 
2 J  inches,  and  for  the  diameter,  2J  inches.  By  the  term  "  length,"  we  mean  only  that 
portion  of  the  pin  which  is  covered  by  the  bushing,  or  the  brass  bearing,  in  which  it 
works.  An  increase  in  the  steam  pressure  of  10  pounds  per  square  inch  of  piston 
will  only  slightly  increase  the  dimensions  of  the  pins ;  hence,  in  the  following  tables 
we  have  given  one  set  of  dimensions,  suitable  for  steam  pressures  varying  from  120  to 
140  pounds  per  square  inch  of  piston,  and  another  set  for  steam  pressures  varying 
from  140  to  160  pounds  per  square  inch. 


TABLE  57. 

DIMENSIONS  OP  KNUCKLE-JOINT  PINS  FOR  MOGUL,  TEN-WHEELED,  AND  CONSOLIDATION  ENGINES,  SUIT- 
ABLE FOR  MAXIMUM  STEAM  PRESSURES  ON  PISTONS  VARYING  FROM  120  TO  140  POUNDS  PER 
SQUARE  INCH. 


Diameter  of  Cylinders. 

Diameter  of  Pins. 

Length  of  Pint. 

11 

H  inches. 

1^  ilK-llfH. 

12 

li 

li 

13 

li 

1ft 

14 

2 

1  ,"„ 

15 

24 

Hi 

16 

2i 

li 

17 

2f 

li 

18 

•J. 

2 

l!i 

2| 

24 

20 

2i 

2* 

21 

3 

2f 

22 

34 

»i 

342 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


TABLE   58. 

DIMENSIONS  OF  KNUCKLE-JOINT  PINS  FOB  MOGUL,  TEN-WHEELED,  AND  CONSOLIDATION  ENGINES,  SUIT- 
ABLE FOB  MAXIMUM  STEAM  PBESSUBES  ON  PISTONS  VARYING  FBOM  140  TO  160  POUNDS  PEE 
SQUABE  INCH. 


Diameter  of  Cylinders. 

Diameter  of  Pins. 

Length  of  Pins. 

11 

If  inches. 

IT*,;  inches. 

12 

U 

1ft 

13 
14 
15 

2 
2* 
2* 

1  ,'•',; 

1 

16 

2f 

1{£ 

17 

if 

2 

18 

2$ 

2i 

19 

3 

2J 

20 

3 

2| 

21 

3i 

2* 

22 

3i 

2f 

CHAPTER  VIII. 

THROTTLE    PIPES.— THROTTLE    VALVE    GEAR.— SAFETY    VALVES.— WHISTLE.— 

PUMPS.— CHECK  VALVES. 

THROTTLE   VALVES. 

349.  Throttle  valves  are  occasionally  placed  in  the  smoke-box,  close  to  the  front 
flue  sheet ;  when  placed  there,  the  throttle  valve  is  simply  a  plain  slide,  arranged  so  as 
to  open  or  close  rectangular  ports,  which  lead  the  steam  into  the  steam  pipes ;  but  the 
general  practice  is  to  place  the  throttle  valve  inside  of  the  dome ;  and  for  such  cases  a 
double  poppet  valve  is  used.  The  design  of  this  valve,  and  that  of  the  throttle  pipe,  is 
shown  in  Figs.  492  and  493.  The  throttle  valve  consists  of  two  disks,  E  E,  cast  to 
three  or  four  wings,  F  F.  The  upper  portion  of  each  one  of  these  disks  is  generally 
bounded  by  a  cylindrical  surface,  and  the  lower  portion,  by  a  conical  surface ;  the 
conical  portions  of  the  disks  fit  into  seats  of  corresponding  form  in  the  throttle  pipe. 
The  angle  formed  by  an  element  of  this  conical  surface  and  the  axis  of  the  valve  is 
often  equal  to  45  degrees,  but  sometimes  it  is  less,  as  shown  in  our  illustrations. 
Necessity  demands  a  greater  diameter  for  the  upper  disk  than  for  that  of  the  lower  one, 
so  as  to  enable  us  to  pass  the  lower  valve  through  the  upper  opening.  A  difference  in 
the  diameters  of  these  disks  is  in  nowise  objectionable;  in  fact,  it  is  desirable  and 
advantageous  for  the  following  reason :  the  throttle  pipe  is  surrounded  by  the  steam 
in  the  boiler,  and  when  the  valve  is  closed  a  greater  pressure  will  be  on  the  upper  disk 
than  on  the  lower  one,  because  the  former  exposes  a  larger  area  to  the  steam  pressure 
than  the  latter.  Under  these  conditions  the  valve  has  a  tendency  to  remain  closed 
when  steam  is  shut  off,  which  is  of  great  importance,  as  this  reduces  the  liability  of 
engines  running  away,  thus  sometimes  avoiding  serious  accidents. 

The  throttle  valve  is  placed  in  a  vertical  position,  and  as  high  in  the  dome  as 
possible,  leaving  between  it  and  the  dome-cover  sufficient  room  only  for  raising  the 
valve  so  as  to  obtain  around  it  the  necessary  amount  of  opening  for  the  admission  of 
steam.  The  vertical  distance  through  which  the  valve  must  be  raised  rarely  exceeds 
1 J  inches ;  often  it  is  less. 

The  valve  is  made  to  fit  the  valve-stem  G  quite  easy,  so  as  to  give  the  valve 
ample  freedom  to  come  in  contact  with  its  seat  throughout.  The  stem  G  connects  to 
the  bell  crank  /?,  and  the  latter  connects  to  the  valve-rod  H,  which  passes  through  the 
end  of  the  boiler  and  connects  to  the  throttle  lever.  The  valve-stem  G  is  always  made 
of  wrought-iron ;  the  bell  crank  B  is  sometimes  made  of  cast-iron,  frequently  of 
wrought-iron.  The  valve-rod  -fiTis  usually  a  plain,  round  wrought-irou  bar,  with  brass, 
cast-  or  wrought-iron  ends  I  screwed  on  to  it. 


344 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


The  throttle  pipe,  sometimes  called  the  stand  pipe,  is  occasionally  made  in  two 
pieces,  P  and  (7,  as  shown  in  Fig.  493 ;  and  in  many  engines  it  is  made  in  one  piece. 
When  made  in  two  pieces,  the  joint  between  the  flanges  M  M  should  be  a  ground  one. 
This  design  of  throttle  pipe  (that  is,  one  made  in  two  pieces)  is  generally  used  for  a 
dome,  having  a  portion  of  its  top  riveted  to  the  dome  sheet,  leaving  a  comparatively 
small  opening  for  a  man  to  enter  the  dome  for  the  purpose  of  making  the  connections 


When  Throttle-valve  i*\ 
cloned  the  lever  B  < hould  j- 
touch  the  projection  A,  \ 
on  Standpipe  C;  Then 


VII     .11  II  III!  Ill  I"      W|       I  II'  II 

the  Valve  must  have  Vis 
play  at  />. 


between  the  throttle  and  dry  pipes,  that  between  the  bell  crank  and  throttle  rod,  and 
others.  In  a  dome  with  a  small  opening  at  top  these  connections  cannot  be  con- 
veniently made,  and  in  some  cases  it  is  impossible  to  make  them,  with  the  throttle 
valve  in  position ;  hence  the  pipe  is  made  in  two  pieces,  so  that  everything  can  be 
properly  and  securely  connected,  before  the  upper  portion  P,  containing  the  throttle 
valve,  is  placed  in  position.  For  domes  whose  whole  top  can  be  removed,  throttle 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


345 


pipes  made  in  one  piece  are  suitable.  Throttle  pipes  are  always  made  of  cast-iron,  and 
therefore  the  throttle  valves  should  also  be  made  of  the  same  nietal,  so  as  to  obtain  an 
equal  rate  of  expansion  in  the  valve  and  the  pipe,  thereby  preventing  leakage.  It  is 
on  account  of  the  difference  in  the  rate  of  expansion  of  the  different  metals  that  brass 
valves  in  cast-iron  pipes  have  proved  to  be  a  failure.  Even  cast-iron  valves  will 
expand  lengthways  a  trifle  more  than  the  pipe,  and  this  fact  should  not  be  overlooked 
in  fitting  the  valve  to  its  seat.  In  order  to  obtain  a  steam-tight  joint,  the  valve  is 
ground  on  its  seats,  until  a  perfect  fit  between  both  the  tipper  and  lower  disks  and  their 


3  Wtngt 
(thick 

Fig.  49C 


T 

DJ 

u 

, 

* 

3  Wtngi 
i  thick 


Fig., 407 


14ft  of  Valve  t 


Fly.  494 


Fig.,498 


Fig.  499 


seats  has  been  obtained ;  the  emery  should  then  be  wiped  off  the  upper  disk  and  its 
seat,  and  a  few  extra  turns  given  to  the  valves,  so  as  to  very  slightly  ease  the  fit 
between  its  lower  disk  and  seat ;  fitting  the  valve  in  this  manner  will  generally  secure 
a  steam-tight  throttle  when  the  engine  is  under  steam. 

350.  Figs.  494,  495  represent  another  design  of  throttle  pipe;  Fig.  496  repre- 
sents the  valve.  The  principal  difference  between  the  throttle  pipe  shown  in  Fig. 
493  and  the  one  shown  here  is  that  the  latter  has  at  the  top  a  seat  A  for  a  snmll 
relief  valve,  which  is  shown  in  Fig.  497.  The  object  of  this  relief  valve  is  to  prevent 


346 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


the  dry  pipe  N  N  (Fig.  500)  from  bursting,  which,  without  this  valve,  is  liable  to  occur 
when  the  engine  is  suddenly  reversed.  Springs  are  not  used  for  holding  this  valve  on 
its  seat ;  the  steam  pressure  in  the  boiler  is  sufficient  to  do  so.  When  the  valve  does 
lift,  it  is  prevented  from  lifting  too  high  by  the  small  wrought-iron  plate  B,  held  in 
position  by  the  two  studs  C  C. 

The  elbow,  Figs.  498, 499,  is  bolted  to  the  throttle  pipe ;  the  joint  E  is  a  ground  joint. 

The  dry  pipe  N  N  is  made  of  wrought-iron ;  the  thickness  varies  for  different 
sizes  of  pipes  from  7^  to  f$  inch ;  the  former  thickness  is  suitable  for  pipes  4  inches 
outside  diameter.  But  dry  pipes  of  the  same  diameter  will  not  always  have  the  same 
thickness,  as  master-mechanics  will  increase  it  to  suit  the  steam  pressure  in  the  boiler ; 
the  thickness  is  also  frequently  increased  for  pipes  longer  than  the  average  length  of 
pipes.  Dry  pipes  1  inches  outside  diameter  are  the  largest  that  we  have  seen 
used.  We  believe  that  for  this  size  of  pipe  a  thickness  of  -fg  inch  is  sufficient 
for  the  greatest  length  required  in  any  locomotive.  The  smallest  outside  diameter  of 
a  dry  pipe  that  we  have  seen  used  is  4  inches;  the  diameter  for  the  next  size  is 
4 \  inches ;  diameters  larger  than  5  inches  increase  by  1  inch.  Dry  pipes,  like  boiler 
tubes,  are  designated  by  their  outside  diameter. 

351.  A  sleeve,  generally  made  of  brass,  is  attached  to  each  end  of  the  dry  pipe. 
In  some  cases  the  dry  pipe  is  slipped  over  the  sleeves,  but  frequently  the  sleeves  /  and 
J  are  slipped  over  the  dry  pipe,  as  shown  in  Fig.  500. 

When  the  diameter  determined  by  the  rule  which  we  shall  presently  give  agrees 


Fig-  BOO 


with  one  of  the  standard  diameters  of  pipes,  the  best  practice  is  to  put  the  dry  pipe 
inside  of  the  sleeves. 

Some  master-mechanics  shrink  the  sleeves  on  to  the  pipe;  others  fasten  them 
with  rivets,  generally  f  inch  diameter.  Two  rows  of  rivets,  from  l£  to  2  inches 
apart,  are  used.  The  pitch  of  rivets  is  generally  from  2  to  3  inches.  The  rivets  are 
arranged  zigzag ;  in  some  instances  twice  as  many  rivets  in  one  row  as  in  the  other 
are  used,  as  shown  in  Fig.  500.  Frequently,  instead  of  using  rivets,  the  holes  through 


MODERN  LOCOMOTim   CONSTRUCTION. 


347 


the  sleeve  and  dry  pipe  are  tapped,  threaded  plugs  screwed  in,  and  their  ends  slightly 
riveted  over,  to  prevent  them  from  turning  and  falling  out.  In  our  opinion,  the  best 
practice  is  to  use  iron  rivets  or  plugs ;  copper  rivets  should  be  avoided,  as  the  action 
of  some  kinds  of  water  will  produce  corrosion  between  the  iron  and  copper,  and  con- 
sequently cause  leakage. 

When  the  sleeves  are  placed  on  the  outside  of  the  pipe,  and  fastened  with  rivets 
or  plugs,  as  mentioned,  they  are  calked ;  or  if  the  shape  of  the  sleeve  will  allow  it, 
the  ends  of  the  pipe  are  calked.  Pipes  placed  on  the  outside  of  the  sleeves  always 
have  their  ends  calked,  but  sleeves  shrunk  on  the  pipe  are  frequently  not  calked. 
Sometimes  the  ends  of  the  pipe  are  threaded,  and  the  sleeves  screwed  on ;  in  such 
cases,  calking  will  be  a  detriment. 

352.  The  end  G  of  the  sleeve  I  (Fig.  500)  is  turned  to  a  spherical  form,  and  the 
elbow  at  F  (Fig.  498)  is  counterbored  to  a  similar  shape,  so  as  to  make  a  ball  joint  between 


Tig.  B01 


Wrought  Iron 


e -jtf 4, 


^  —      -] 

•> 

1  '' 

kid 

i 

^ 

in 

y 

+ 

the  two.  For  connecting  the  diy  pipe  to  the  elbow,  the  yoke  shown  in  Figs.  501,  502, 
503  is  slipped  over  the  sleeve  /and  the  elbow  F;  the  lugs  Y  Y  are  made  to  bear  against 
the  flat  end  of  the  sleeve,  the  point  K  of  the  set  screw  is  inserted  in  the  countersunk 
cavity  L  in  the  lug  M  (Fig.  498),  and  the  sleeve  I  drawn  tightly  against  spherical  face  F, 
and  fii'mly  held  there.  Although  this  mode  of  fastening  the  dry  pipe  to  throttle  pipe  is 
often  used,  and  gives  good  satisfaction,  other  designs  for  accomplishing  the  same  thing 
are  adopted ;  for  instance,  for  a  throttle  pipe  like  that  shown  in  Fig.  493,  two  hooked 
bolts  instead  of  a  yoke  are  used.  But  in  all  cases  the  joint  between  the  throttle  pipe 
and  dry  pipe  is  a  ball  joint,  which  affords  ready  means  for  adjusting  the  dry  pipe  to 
any  inaccuracies  which  cannot  be  avoided  in  the  construction  of  a  boiler.  The  sleeve 
J  at  the  other  end  of  the  dry  pipe  (Fig.  500)  is  also  made  so  as  to  form  a  ball  joint,  with 
a  casting  riveted  to  the  front  flue  sheet ;  this  casting  or  ring  will  be  presently  shown. 

353.  Figs.  504,  505  represent  another  throttle  pipe,  and  Figs.  506,  507,  508  show 
some  of  its  details.  The  design  of  this  throttle  pipe  does  not  differ  much  from  those 
previously  shown.  The  principal  differences  ai-e  that  the  pipe  is  made  in  one 
piece,  and  that  the  sleeve  7  is  of  a  different  form.  The  manner  of  fastening  the 


348 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


dry  pipe   to   the   throttle  pipe  is  plainly  shown,   and  does  not  need  any  further 
explanation. 

The  favorite  way  of  fastening  the  throttle  pipe  to  the  dome  is  plainly  shown  in 


• 4) 


p 

• 

^ 

• 

V 

7) 

Flg.i 

i 

' 

p_ 
1 

i 

i 

_._ 

„•-.- 

r 

1 

-* 

! 

F 
J            w 

aW/ttWWy           " 

[f 

1 

; 
- 

ijj 

r'1 

i    i   V 

'H  f  I 

^! 
fl'f 

.. 

5    -- 

j  -: 

!    

3^ 

!^"~* 

c-i-4- 

-[4 
-ii  —  *-• 

i 
i  ,m  [ 

•f                                   1 

A"    ! 

I 

! 

if 

sffi 

*I  * 

^^® 

r 
i 
i 

a,    y. 
f    f           f 

Jlij/:.  S04 


r 


these  figures.     A  bracket  D  is  cast  to  the  throttle  pipe,  and  fastened  to  the  dome  sheet 
F  by  two  studs.     When  possible,  all  dry  pipes  are  allowed  to  rest  on  top  of  the  crown 


MODERN  LOCOMOTIVE  CONSTRUCTION.  349 

bars.  When  this  cannot  be  done,  the  dry  pipe  is  suspended  by  a  strap,  usually  made 
of  -J  x  i-inch  iron,  bolted  to  the  boiler  shell  W. 

The  thickness  of  the  throttle  pipes  is  generally  about  £  to  f  inch. 

Fig.  504  represents  a  pipe  exceptionally  thin,  the  thickness  being  only  -fa  of  an 
inch ;  and  for  this  reason  the  thickness  at  the  lower  end  is  increased,  so  as  not  to 
injure  it  in  clamping  the  dry  pipe  to  it. 

The  inner  diameter  of  the  throttle  pipe  is  generally  made  proportionate  to  the 
diameter  of  the  cylinder ;  the  ratio  between  the  two,  as  made  by  different  builders, 
varies.  Hence  we  sometimes  find  the  inner  diameter  of  the  throttle  pipe  equal  to  one- 
quarter  of  the  diameter  of  the  cylinder,  and  frequently  we  find  it  to  be  equal  to  about 
one-third  of  the  diameter  of  the  cylinder.  The  latter  we  believe  to  be  the  best  propor- 
tion, and  should  be  adopted.  When  the  inner  diameter  of  the  throttle  pipe  is  not 
uniform  throughout — for  instance,  if  the  throttle  pipe  is  made  similar  to  that  shown  in 
Fig.  504 — the  smaller  diameter  P  should  be  one-third  of  the  diameter  of  the  cylinder. 
The  inner  diameter  of  a  throttle  pipe  should  not  be  greater,  in  any  case,  than  here 
given.  If  made  much  smaller,  so  that  the  cross-sectional  area  of  the  throttle  pipe 
is  less  than  one-tenth  of  the  cross-sectional  area  of  the  cylinder,  the  initial  steam 
pressure  in  the  cylinder  at  high  speeds  will  be  reduced  below  the  boiler  pressure 
more  than  it  would  be  otherwise,  and  the  tractive  power  of  the  engine  will  be 
interfered  with.  On  the  other  hand,  practice  seems  to  indicate  that  no  advantage  is 
gained  by  making  the  inner  diameter  of  the  throttle  pipe  greater  than  one-third  of  the 
diameter  of  the  cylinder ;  in  fact,  throttle  pipes  and  diy  pipes  too  large  in  diameter  are 
detriments  to  the  engines,  because  an  unnecessary  quantity  of  steam  held  in  these 
pipes  must  be  worked  off  before  the  engine  can  be  stopped,  which  is  an  objection  in 
case  of  an  emergency. 

The  cross-sectional  area  at  0,  Fig.  504,  is  generally  rectangular  in  form ;  the 
length  and  breadth  of  this  section  should  be  so  proportioned  as  to  give  an  area  equal 
to  that  through  the  cylindrical  portion  of  the  pipe. 

354.  The  inner  diameter  of  the  dry  pipe  should  be  equal  to  that  of  the  throttle 
pipe;  but  since  in  many  cases  the  diameter  of  the  pipe,  determined  by  calculation, 
cannot  be  found  among  the  standard  diameters  of  dry  pipes,  the  next  larger  size  is 
adopted ;  and  therefore  we  often  find  locomotives  having  dry  pipes  whose  inner  diame- 
ters are  greater  than  those  of  the  throttle  pipes. 

The  outer  diameter  of  the  upper  disk  of  the  throttle  valve  is  sometimes  a  little 
greater  than  the  inner  diameter  of  the  throttle  pipe ;  frequently  it  is  less.  A  good 
rule  is  to  make  the  outer  diameter  of  the  upper  disk  equal  to  one-third  of  the  diameter 
of  the  cylinder.  The  outer  diameter  of  the  lower  disk  is  made  from  £  to  £  of  an  inch 
less  than  the  diameter  of  the  upper  opening  in  the  valve  seat. 

355.  Fig.  510  represents  a  complete  sectional  view  of  the  connections  of  the 
T-pipe  and  dry  pipe  N.    The  brass  ring  Bz  is  riveted  to  the  flue  sheet  A ;  another 
view  of  this  ring  is  shown  in  Fig.  509.     The  six  holes  marked  s  s  in  Fig.  509  are 
tapped  for  the  studs  marked  F  F  in  Fig.  510 ;  the  remaining  holes  r  r  are  the  rivet 
holes.     The  size  of  rivets  and  studs  vary  according  to  the  size  of  engine ;  for  small 
engines,  say  with  cylinders  10  inches  diameter,  the  rivets  through  the  ring  11.,  are 
generally  f  inch  diameter,  and  the  studs  f  inch  diameter ;  for  engines  with  cylinders 


350 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


20  inches  diameter,  rivets  3  inch  di- 
ameter and  studs  1  inch  diameter  are 
generally  used.  The  flue  sheet  A  is 
bored  out  to  receive  the  projection  on 
the  brass  ring  B2.  The  spherical  seat 
in  this  ring  is  made  to  fit  the  spherical 
surface  on  the  sleeve  J\  this  sleeve,  we 
have  seen  in  Art.  351,  is  riveted  to 
the  dry  pipe  2V.  The  T-pipe  has  a 
spherical  projection  p,  bearing  against 
the  surface  of  the  counterbore  in  the 
sleeve  J ;  some  builders  make  this 
couuterbore  a  conical  surface;  others 
make  it  a  spherical  surface  as  shown ; 
but  in  all  cases,  the  projection  p  on  the 
the  T-pipe  is  turned  to  a  spherical  form. 

The  diameter  of  the  opening  M  in 
the  brass  ring  must  be  made  suffi- 
ciently large  to  admit  the  sleeve  on 
the  other  end  of  the  dry  pipe;  this 
condition  will  determine  to  a  great  ex- 
tent the  size  of  the  ball  joint  on  the 
sleeve  J.  It  will  be  seen  that  the 
studs  F  F  force  the  T-pipe  against 
the  sleeve  J,  and  the  latter  is,  in  turn, 
pressed  against  the  ring  B2,  thus  form- 
ing steam-tight  joints  between  the 
whole. 

Some  builders  do  not  use  a  brass 
ring  B2  for  small  locomotives,  but  use 
in  place  of  it  a  wrought-iron  plate  P  P, 
as  shown  in  Fig.  512.  In  fact,  some 
master-mechanics  prefer  the  wrought- 
iron  plate  P  P  for  all  sizes  of  engines, 
and  consequently  we  frequently  meet 
with  large  engines  which  have  a  plate 
of  this  kind  in  place  of  the  brass  ring. 
The  plate  is  riveted  to  the  flue  sheet 
A  A ;  the  arrangement  of  rivets  and 
studs  for  holding  the  T-pipe  is  similar 
to  that  in  the  brass  ring  as  shown  in 
Pig.  509.  Both  the  flue  sheet  and  the 
plate  are  then  couuterbored  as  shown, 
so  as  to  form  a  bearing  for  the  spher- 
ical part  of  the  sleeve  J. 


MOVER*  LOCOMOTIVE   CONSTRUCTION. 


351 


356.  The  right-hand  side  of  Fig.  511  represents  an  outside  view  of  the  T-pipe, 
and  the  left-hand  side  represents  a  section  of  the  connection  of  the  T-pipe  and 
the  steam  pipe  1).     The  steam  pipes  lead  to  the  cylinders.     A  complete  drawing  of 
these  pipes  and  their  position  in  the  smoke-box  will  be  found  in  Fig.  24.     In  fact  the 

.illustrations  here  given  simply  show  in  detail  and  to  a  larger  scale  the  connections  of 
the  dry  pipe,  T-pipe,  and  steam  pipes. 

The  opening  in  the  flanges  E  E  of  the  T-pipe  are  counterbored  to  a  spherical  form 
for  the  brass  rings  C,  which  are  inserted  between  the  T-pipe  and  steam  pipes.  Usually 
two  bolts  G  G  are  used  for  connecting  each  steam  pipe  to  the  T-pipe. 

The  inside  diameter  K  of  the  T-pipe  must  be  equal  to  that  of  the  throttle  pipe. 
The  area  of  the  opening  L  in  the  branches  of  the  T-pipe  should  be  equal  to  the  inner 
cross-sectional  area  of  the  stearn  pipe.  The  rale  for  finding  this  area  has  been  given 
in  Art.  46. 

357.  Fig.  511A  shows  a  throttle  pipe  arranged  to  take  a  device  for  furnishing  dry 
steam  to  the  cylinders.    This  is  accomplished  by  separating  the  steam  from  the  water 


Fig.  5 11 A 


when  the  engine  is  running.  The  device  is  called  a  separator,  and  consists  of  a  single 
casting  with  the  necessary  drain  pipes  for  leading  off  the  water  after  it  has  been 
separated  from  the  steam.  It  will  presently  be  seen  that  its  construction  is  exceedingly 
simple ;  it  is  very  durable,  and  requires  veiy  little  or  no  attention. 

The  separator  is  cast  in  one  piece.  Its  core  B  is  made  hollow,  and  is  gradually 
reduced  from  a  comparatively  large  diameter  at  the  center  to  a  point  at  each  end,  so  as 
to  form  a  conoidal  surface.  A  number  of  wings  C  C  are  cast  to  this  surface  and 


352  MODERN  LOCOMOTIVE   CONSTRUCTION. 

extend  spirally  towards  the  ends  of  the  core.  This  separator  is  set  concentric  with 
the  throttle  pipe,  whose  diameter  is  necessarily  somewhat  larger  than  that  of  an 
ordinary  one.  The  separator  does  not  rotate,  but  it  is  firmly  attached  to  the  pipe  as 
shown. 

In  Fig.  511B  it  will  be  noticed  that  one  side  of  each  wing  is  formed  tangential  to 
the  surface  of  the  core,  and  the  other  side  approaches  a  radial  surface.  The  object 
of  the  whole  construction  is  to  divide  the  steam  as  it  flows  through  the  pipe  into 
several  smaller  currents,  and  to  give  to  each  a  compound  whirling  motion  which  is 
accomplished  in  the  following  manner.  The  core  B  of  the  separator  spreads  the  steam 
or,  so  to  speak,  expands  it  into  an  annular  body,  and  the  wings  C  C  divide  it  into 
several  streams  or  currents,  while  their  spiral  forms  impart  to  each  current  a  whirling 
motion  around  the  core ;  and  the  sides  tangential  to  the  surface  of  the  core  impart  to 
each  stream  a  whirling  motion  within  itself.  It  will  also  be  noticed  that  by  expanding 
the  steam  into  an  annular  body  and  cutting  it  up  into  several  streams  all  the  suspended 
particles  of  water  will  be  affected  by  the  whirling  motion. 

If,  on  the  other  hand,  the  solid  stream  as  it  enters  the  separator  had  not  been 
expanded  into  an  annular  body,  the  whirling  motion  could  not  act  on  the  particles  of 
water  in  the  center,  and  the  action  of  the  separator  would  thereby  be  impaired.  The 
result  of  all  this  is  that  as  the  steam  enters  the  throttle  pipe  a  violent,  compound 
whirling  motion  is  imparted  to  all  the  particles  of  steam  and  water,  all  the  heavier 
particles  are  thrown  against  the  pipe,  and  the  water  thus  separated  from  the  steam 
flows  into  the  annular  chamber  at  the  bottom  of  the  throttle  pipe,  whence  it  is  con- 
ducted through  the  tube  J  to  the  outside  of  the  boiler  and  may  be  fed  back  to  it  if 
desirable.  The  dry  steam  is  conducted  through  the  vertical  branch  H  into  the  dry 
pipe  leading  to  the  cylinders.  The  bell-shaped  cup  G  gathers  any  water  which  may 
flow  along  the  core  B,  which  is  discharged  into  the  tube  7,  which  carries  it  off  with  the 
rest  of  the  water. 

In  applying  this  separator  to  a  locomotive  nothing  needs  to  be  changed  excepting 
the  throttle  pipe.  There  is  nothing  to  get  out  of  order,  which  is  an  important  feature 
in  mechanical  devices  which  are  placed  out  of  sight  and  cannot  be  reached  like  a  throttle 
pipe  in  a  locomotive  boiler.  This  separator  is  the  invention  of  Mr.  Joseph  De  Rycke 
of  New  York.  It  has  been  successfully  used  in  many  marine  engines  and  on  steam 
mains  from  200  to  800  feet  in  length ;  but  we  are  not  aware  that  it  has  yet  been  applied 
to  locomotives,  for  which  we  believe  it  is  well  adapted. 

THROTTLE  VALVE   CONNECTIONS. 

358.  There  are  various  ways  of  attaching  the  throttle  valve  connections  to  the 
boiler.  Sometimes  we  are  compelled  to  run  the  throttle  lever  connections  through  the 
top  of  the  boiler ;  but  generally  the  throttle  rod  passes  either  through  the  back  head 
of  the  boiler,  or  it  passes  through  the  sides  of  the  dome.  Figs.  513,  514,  515  repre- 
sent a  throttle  valve  lever  and  its  attachment,  suitable  for  engines  in  which  the 
throttle  rod  H  is  passed  through  the  back  head  of  the  boiler. 

The  throttle  lever  is  probably  the  simplest  in  design  used  on  locomotives.  The 
back  head  P  P  of  the  boiler  is  bored  out  to  receive  the  spherical  portion  of  the  stuffing- 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


353 


box  C,  the  two  forming  a  ball  joint ;  this  joint  is  a  ground  one,  so  as  to  make  it 
perfectly  steam  tight.  The  stuffing-box  is  fastened  to  the  boiler  head  by  four  studs, 
.L,  M,  K,  K;  the  studs  K,  K  are  made  long  enough  to  take  hold  of  the  stuffing-box 
gland.  The  form  of  the  nut  on  the  stud  M  is  made  suitable  for  receiving  a  pin,  on 
which  the  two  links  0,  0  vibrate.  These  links  are  connected  to  the  throttle  lever 
A,  and  act  as  a  fulcrum.  The  throttle  lever  A  is  connected  to  the  throttle  rod  H  by 
means  of  the  jaw  7.  In  a  few  instances  this  jaw  is  screwed  on  to  the  rod  H,  but 


generally  the  jaw  is  bored  out  to  a  taper,  accurately  fitting  the  tapered  end  of  the  rod 
H,  which  is  driven  into  the  jaw,  and  held  there  by  means  of  the  tapered  pins  G,  G ;  in 
some  cases  only  one  pin  in  place  of  two  is  used.  The  taper  on  the  end  of  the  rod  H 
is  generally  £  inch  in  2  inches.  The  throttle  lever  rests  on  the  quadrant  B,  which 
is  fastened  to  the  boiler  head  by  means  of  one  stud;  in  a  few  instances  two 
studs  are  used  for  the  same  purpose.  Through  the  quadrant  B,  a  slot  is  cut  for  the 
clamping  bolt  F;  the  nut  E  for  this  bolt  is  capped  with  wood.  The  throttle  lever  A 
is  clamped  to  the  quadrant  B  in  any  position  which  gives  the  throttle  valve  in  the 
dome  the  desired  degree  of  opening. 

This  design  of  lever  is  often  adopted  on  account  of  its  simplicity ;  but  it  is  not 
a  convenient  one  for  the  engineer  to  handle,  because  in  order  to  open,  close,  or  adjust 
the  throttle  valve  with  this  kind  of  lever  the  engineer  may  have  to  use  both  hands, 
which  is  not  only  inconvenient,  but  in  case  of  an  emergency  is  objectionable. 


354 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


359.  Figs.  516,  517  represent  a  throttle  lever  and  attachments  designed  to  remove 
the  objectionable  feature  inherent  in  the  former  one.  The  principal  difference  between 
the  two  designs  lies  in  the  de- 
sign of  the  quadrant  B. 

In  Fig.  516  it  will  be  seen 
that  on  the  convex  edge  of  the 
quadrant  B  notches  are  cut  which 
engage  with  corresponding  teeth 
on  the  end  of  the  latch  S;  this 
latch  is  connected  to  the  handle 
B  by  the  link  T.  It  will  readily 
be  perceived  that,  with  this 
arrangement,  the  engineer  can 


/•k  ' 

f 

t 

-v    ' 

(          i-H 

:i 

• 

'  i  '      -J, 
!  !  '    1    f 

l-ll  '- 

1 

!  i;      —    -  i! 

r 

UJ 

with  one  hand  disengage  the  latch,  move  the  lever  into  any  position,  and  allow  it  to 
lock  itself  there. 

The  manner  of  fastening  the  quadrant  B  (Fig.  516)  to  the  boiler  differs  somewhat 
from  that  shown  in  Fig.  514,     In  the  former  figure  it  will  be  noticed  that  a  bracket  -B2 


MODERN  LOCOMOTIVE  CONSTRUCTION.  355 

is  fastened  to  the  boiler,  and  the  quadrant  B  is  attached  to  this  bracket.  In  many 
engines  the  bolt  U  does  not  hold  the  quadrant  B  i-igidly  to  the  bracket,  but  allows 
it  to  vibrate  a  little,  so  as  to  adjust  itself  to  any  position  of  the  lever  A.  Under  these 
conditions,  we  need  in  the  quadrant  B  a  slot  cut  equidistant  from  the  notched  edge  of 
the  arc.  In  this  slot  a  bolt  F  is  accurately  fitted,  and  prevents  a  disengagement  of 
the  quadrant  B  and  the  latch  S.  These  are  usually  placed  above  the  lever ;  in  such 
cases,  the  bolt  F  will  tend  to  prevent  the  lever  from  moving  out  of  its  appointed  plane 
of  action,  and  help  to  produce  steadiness  of  motion. 

360.  The  diameter  of  the  throttle  rod  H  varies  for  the  different  sizes  of  engines ; 
for  small  engines  it  is  about  J  inch,  and  for  large  engines  about  Ij  inches. 

To  place  the  throttle  rod  H  in  position,  it  must  be  passed  through  the  stuffing- 
box;  hence,  collars  on  the  i-od  are  not  admissible.  When  these  rods  are  made  of 
uniform  diameter  throughout,  they  must  be  turned  throughout  their  whole  length ;  to 
save  time  in  turning,  a  largo  portion  of  the  rod,  extending  to  within  a  short  distance 
from  the  ends,  is  forged  about  j-6  of  an  inch  smaller  in  diameter  than  the  required 
finished  diameter  at  the  ends,  leaving  at  the  throttle-lever  end  a  portion  to  be  finished, 
of  a  length  only  as  may  be  required  for  the  movement  of  the  rod  in  the  stuffing- 
box,  and  leaving  at  the  throttle-pipe  end  a  portion  of  such  a  length  as  may  be  required 
for  the  thread. 

Sometimes  that  part  of  the  rod  which  works  in  the  stuffing-box  is  provided  with 
a  brass  casing,  as  shown  in  Fig.  516.  This  brass  casing  is  cast  on  the  rod ;  its  object 
is  to  prevent  a  collection  of  rust  between  the  rod  and  stuffing-box.  The  portion  of 
the  rod  covered  by  the  brass  casing  is  generally  forged  to  an  octagon  form. 

361.  The  stuffing-box  is  generally  made   of  cast-iron,  sometimes  of  brass.     For 
large  engines  the  stuffing-box  gland  is  often  made  of  cast-iron.     Glands  made  of  cast- 
iron  have  always  a  brass  bushing.     The  bushing  is  sometimes  driven  into  the  gland 
with  a  hammer,  and  sometimes  it  is  pressed  in  by  a  hydraulic  press  with  a  pressure  of 
about  100  pounds. 

For  small  engines,  the  gland  is  often  made  of  brass.  The  principal  propor- 
tions of  the  gland  and  stuffing-box  are  found  by  the  rule  given  in  Arts.  192,  193,  194, 
and  195. 

362.  The  quadrants  B,  having  a  form  similar  to  that  shown  in  Fig.  514,  are  made 
of  brass,  and  when  they  are  made  like  that  shown  in  Fig.  516  they  should  be  made  of 
the  best  hammered  iron  or  steel,  so  as  to  prevent  wear  of  the  notches.    The  shape  of 
the  notches  is  similar  to  that  of  the  teeth  in  an  ordinary  circular  saw ;  the  notches 
are  cut  so  as  to  bring  their  radial  sides  towards  the  boiler.     Notches  of  this  kind  will 
prevent  the  throttle  valve  from  flying  open,  and  allow  it  to  be  veiy  easily  closed. 

363.  Sometimes  the  throttle  lever  A  is  placed  in  a  horizontal  position ;  in  many 
engines  it  points  upward,  and  occasionally  it  points  downward. 

The  proper  position  of  the  throttle  lever  will  depend  on  the  position  of  the 
stuffing-box  in  the  back  head,  and  the  position  of  the  reverse  lever. 

The  stuffing-box  is  placed  as  high  in  the  back  head  as  is  possible  and  practical  to 
do,  and  determines  the  position  of  that  part  of  the  throttle  lever  which  is  connected  to 
tin'  throttle  rod.  The  handle  of  the  throttle  lever  should  be  as  close  as  possible  to  the 
handle  of  the  reverse  lever,  the  latter  being  placed  in  a  convenient  position  for  the 


356 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


engineer  to  reach;  hence  in  this  way  the  position  of  the  end  of  the  throttle  lever 
is  found,  and  the  plane  in  which  it  is  to  move  is  established. 

The  distance  between  the  extremity  of  the  throttle-lever  handle  and  that  of  the 
reverse  lever  should  be  about  2j  to  3  inches,  so  as  to  prevent  the  engineer  from 
jamming  his  hand  between  the  two ;  hence  this  condition  will  determine  the  length 
of  that  part  of  the  throttle  lever  which  extends  from  the  center  of  the  throttle  rod  to 
the  extremity  of  the  handle. 

364.  The  objection  raised  to  the  throttle  lever  shown  in  Figs.  516,  517  is  that  the 
pitch  of  the  teeth  on  the  quadrant  B  will  limit  the  degree  of  opening  of  the  throttle  valve, 


4 


and  in  some  cases  will  not  be  as  close  to  the  requirements  of  the  engine  as  may  be 
desirable.  Again,  on  account  of  the  fine  pitch  of  the  teeth  in  the  quadrant  7?,  strength, 
and  consequently  security,  will  be  impaired ;  and  also  in  some  cases  the  wear  of  the  teeth 
is  objectionable. 

Figs.  518,  519  represent  a  throttle-valve  lever  designed  to  overcome  these  objec- 
tions. In  this  arrangement  a  curved  rack  B  instead  of  a  quadrant  is  introduced ;  the 
rack  is  placed  under  the  lever  A.  This  rack  engages  with  a  small  pinion  C  keyed  to  the 
stud  D ;  to  the  same  stud  is  keyed  another  wheel  E,  larger  in  diameter  than  the  pinion ; 
the  wheel  E  is  an  internal-spur  wheel,  whose  teeth  engage  with  the  link  T,  preventing 
the  stud  D  from  turning,  and  thus  locks  the  throttle  lever  in  the  desired  position. 
The  reason  for  placing  the  teeth  inside  of  the  wheel  E  is  simply  to  obtain  a  pleasing 
effect,  exposing  to  view  plain  and  polished  surfaces,  which  can  be  kept  clean.  The 


MODERX  LOCOMOTIVE  CONSTRUCTION. 


357 


link  T  is  connected  to  the  handle  E,  the  spring  S  holds  the  link  T  in  contact  with 
the  teeth  in  the  wheel  E.  In  pressing  the  handle  R  towards  the  throttle-lever  handle, 
a  disengagement  of  the  link  T  and  the  wheel  E  takes  place,  enabling  the  engineer  to 
move  the  throttle  lever  to  and  fro  with  ease. 

In  this  aiTangernent  the  wear  of  the  teeth  in  the  rack  will  be  less  than  the  wear  of 
the  teeth  in  the  quadrant  shown  in  Fig.  516 ;  also,  because  the  pitch  of  the  teeth  in  the 
rack  and  in  the  wheel  E  is  greater  than  the  pitch  of  the  teeth  in  the  quadrant  B  (Fig.  516), 
the  strength  of  this  arrangement,  and  security,  is  increased.  A  very  close  adjustment 
of  the  throttle  valve  to  the  requirements  of  the  engine  can  also  be  obtained ;  indeed, 
this  closeness  of  adjustment  largely  depends  upon  the  difference  between  the  diame- 
ters of  the  wheel  E  and  the  pinion  C.  For  instance,  suppose  that  the  pinion  C  makes 
just  one  complete  turn  in  lifting  the  throttle  valve  one  inch,  then  the  wheel  E  must  of 
course  also  make  one  complete  turn  during  the  same  time ;  if  now  the  wheel  E  has  42 
teeth,  then  it  must  be  obvious  that  the  throttle  valve  can  be  set  to  -fa  inch ;  and  if  a 
closer  regulation  is  required  without  changing  the  pitch  of  the  teeth  in  the  wheel  E, 
then  we  have  only  to  increase  its  diameter  so  as  to  enable  us  to  increase  the  number  of 
teeth. 

365.  The  method  for  finding  the  length  and  curvature  of  the  pitch  line  for  the  rack 
B  is  the  same  as  that  for  finding  the  length  and  curvature  of  the  quadrants  B  in  Figs. 
514  and  516.  In  order  to  explain  this  method,  we  have  shown  in  Fig.  524  a  portion  of  the 


throttle  lever  A,  the  links  0,  and  a  portion  of  the  quadrant  B.  Before  we  can  determine 
the  length  and  curvature  of  the  quadrant  we  must  know  the  lift  of  the  throttle  valve ; 
hence  the  following  problems  present  themselves :  first,  to  find  the  lift  of  the  throttle 
valve ;  second,  to  find  the  length  of  the  arc  c  n  d ;  third,  to  find  the  radius  of  the  same  arc. 


358  MODERN  LOCOMOTIVE   CONSTRUCTION. 

First,  to  find  the  lift  of  the  throttle  valve.  To  explain  this,  we  shall  refer  to  Fig. 
504.  In  this  figure  we  see  that  the  smallest  inner  diameter  (4f  inches)  of  the  throttle 
pipe  is  at  the  bottom  of  the  pipe ;  the  area  of  a  circle  4|  inches  diameter  is  16.80  square 
inches ;  hence  the  throttle  valve  must  have  a  lift  which  will  allow  such  a  quantity  of 
steam  to  enter  as  can  pass  through  an  area  of  16.80  square  inches.  Again,  we  see  that 
the  valve  seats  are  of  a  conical  form ;  but,  for  the  sake  of  simplicity  in  finding  the  lift, 
the  valve  seats  are  assumed  to  be  flat.  Under  these  conditions  the  lift  of  the  valve 
must  be  such  that,  when  the  circumferences  of  the  inner  edges  of  the  valve  seats  are 
multiplied  by  the  lift,  the  area  thus  obtained  will  be  equal  to  the  smallest  inner  cross- 
sectional  area  of  the  pipe,  which  in  our  case  is  16.80  squai'e  inches. 

In  Fig.  504  we  see  that  the  diameter  of  the  inner  edge  of  the  upper  throttle-valve 
seat  is  4£  inches,  hence  its  circumference  will  be  4.5  x  3.1416  =  14.13+  inches;  the 
diameter  of  the  inner  edge  of  the  lower  valve  seat  is  4  inches,  its  circumference  will  be 
4  x  3.1416  =  12.56+  inches ;  the  sum  of  the  two  circumferences  will  be  14.13  +  12.56 

=  26.69  inches;  hence  in  the  case  before  us  the  lift  will  be  K-T^  =  .62+  inch.     But 

.iO.uy 

this  lift  is  suitable  only  for  flat  valve  seats,  and  will  not  give  us  a  sufficient  opening 
for  conical  valves.  In  nearly  all  locomotives  the  inclination  of  the  valve  seat  varies 
but  little  from  an  angle  of  45  degrees  from  the  axis,  and  therefore  we  will  generally 
obtain  good  results  by  adding  50  per  cent,  to  the  lift  just  found ;  hence  the  lift  for  the 
valve  shown  in  Fig.  504  should  be  .62  +  .31  =  .93  inch.  From  the  foregoing  we  may 
establish  the  following  rule : 

RULE  86. — Divide  the  smallest  cross-sectional  area  of  the  throttle  pipe  by  the  sum 
of  the  circumferences  of  the  openings  in  the  valve  seats ;  add  50  per  cent,  to  the 
quotient ;  the  sum  will  be  the  lift  of  the  valve. 

If  now  the  lengths  of  the  arms  of  the  bell  crank  to  which  the  valve-stem  and 
throttle  rods  are  connected  had  been  equal,  the  throttle  rod  H  would  have  to  move 
through  a  distance  of  .93  inch.  But  in  Fig.  504  we  see  that  the  arm  of  the  bell  crank 
to  which  the  valve-stem  connects  is  2£  inches  long,  and  the  other  arm  is  9  inches 

9 
long,  hence  the  movement  of  the  throttle  rod  will  be  ~  =  3.6  times  greater  than  the 

ay 

lift  of  the  valve ;  and  if  the  lift  of  the  valve  is  .93  inch,  then  the  total  movement  of 
the  throttle  rod  will  be  3.6  x  .93  =  3.348  inches,  say  3f  inches. 

Second,  to  find  the  length  of  the  arc  c  d,  Fig.  524. 

To  make  the  solution  of  our  problem  as  plain  as  possible,  let  us  assume  that 
Fig.  524  is  a  portion  of  the  throttle  work,  as  shown  in  Fig.  514. 

Draw  the  center  line  k  I  of  the  throttle  rod  H,  and  on  it  lay  off  two  points,  k  and  / ; 
the  distance  between  these  points  must  be  equal  to  the  travel  or  movement  of  the 
throttle  rod ;  if  this  travel  is  to  be  3f  inches,  as  found  by  the  foregoing  calculations, 
then  make  the  distance  between  k  and  I  equal  to  3f  inches.  When  the  throttle- valve 
is  one-half  open,  or,  in  other  words,  when  it  stands  in  the  center  of  its  lift,  the  center 
line  p  t  of  the  throttle  lever  A  stands  generally  parallel  to  the  back  end  of  the  boiler 
and  the  center  line  e/of  the  link  0  stands  perpendicular  to  p  t,  as  shown  in  Fig.  514. 
Under  these  conditions,  draw  through  the  center  m  (Fig.  524)  of  the  movement  k  I  a 
t  perpendicular  to  k  /,  and  a  line  cf  perpendicular  to  p  t,  cutting  the  latter  in  the 


MODERN  LOCOMOTIVE   CONSTRUCTION.  359 

point/;  the  distance  between  the  lines  k  I  and  ef  must  of  course  be  equal  to  the  given 
distance  between  the  center  line  of  the  link  0  and  the  center  line  of  the  throttle  rod ; 
in  Fig.  514  we  see  that  this  distance  is  3  inches.  The  point /will  be  the  center  of  the 
fulcrum  pin  through  the  end  of  the  lever  A,  and  the  point  m  will  be  the  center  of  the 
pin  through  the  lever  A  and  the  throttle  rod  H.  On  the  line  p  t  lay  off  a  point  n,  and 
make  the  distance  between  m  and  n  equal  to  the  given  distance  between  the  center  of 
the  pin  m  and  the  center  of  the  clamping  bolt ;  in  Fig.  513  we  see  that  this  distance  is 
6£  inches.  Now,  the  center  e  in  the  link  0  will  be  stationary,  the  center  /will  move 
along  the  arc  g  It  described  from  the  center  e,  and  the  center  m  will  move  along  the 
straight  line  A;  /.  Therefore  from  the  point  A;  as  a  center,  and  with  a  radius  equal  to 
the  distance  between  /  and  m,  describe  a  short  arc  cutting  the  arc  g  h  in  the  point  i, 
join  the  points  i  and  A;  by  a  straight  line,  and  prolong  it  towards  r.  Again,  from  the 
point  /  as  a  center,  and  with  a  radius  equal  to  fm,  describe  a  short  arc  cutting  the 
arc  g  h ;  the  point  in  which  these  two  arcs  intersect  will  very  nearly  coincide  with  the 
point  i,  previously  found ;  for  all  practical  purposes  we  may  assume  that  these  points 
coincide.  Through  the  points  i  and  I  draw  a  straight  line,  and  prolong  it  towards  s. 
Make  k  r  equal  to  m  n ;  also  make  I  s  equal  to  m  n. 

When  the  throttle  valve  is  closed,  the  center  line  of  the  throttle  lever  A  will 
coincide  with  the  line  i  r,  and  the  point  r  will  be  the  position  of  the  center  of  the 
clamping  bolt.  When  the  throttle  valve  is  full  open,  the  center  line  of  the  throttle 
lever  A  will  coincide  with  the  line  i  s,  and  the  point  s  will  be  the  position  of  the  center 
of  the  clamping  bolt.  The  arc  r  n  s  will  represent  the  length  of  the  path  of  the  center 
of  the  clamping  bolt.  In  practice  it  is  customary  to  make  the  arc  r  n  s  about  1  inch 
longer  than  length  just  found,  consequently  the  length  of  the  arc  end  will  be  equal  to 
the  sum  of  the  arc  r  n  s,  as  found  by  construction,  plus  the  diameter  of  the  clamping 
bolt,  plus  1  inch. 

Third,  to  find  the  radius  of  the  arc  end.  This  arc  will  not  coincide  exactly  with 
an  arc  of  a  circle,  yet  the  difference  is  so  slight  that  it  may  be  neglected,  and  for  all 
practical  purposes  we  may  assume  that  the  arc  c  n  d  is  an  arc  of  a  circle.  Now,  the 
points  r  n  s  are  points  in  this  arc,  hence  all  that  is  necessary  is  to  find  a  point  p,  from 
which  an  arc  can  be  described  which  will  pass  through  the  three  points  r  n  s;  the 
distance  from  p  to  any  one  of  these  points  will,  of  course,  be  the  required  radius. 

Probably  the  quickest  way  to  find  the  point  p  is  by  trial ;  it  can  also  be  found  in 
a  geometrical  way,  by  joining  the  points  r  and  n  by  a  straight  line ;  also  joining  the 
points  n  and  s  by  a  straight  line.  Then  bisect  the  lines  r  n  and  n  s  by  perpendicular 
lines;  the  point  in  which  these  perpendiculars  intersect  will  be  the  center  _p  from  which 
the  arc  r  n  s  is  to  be  described. 

The  lines  u  v  and  iv  x  extend,  and  ai-e  perpendicular  to  the  back  head  of  the  boiler ; 
the  distance  between  the  center  line  of  the  throttle  rod  H  and  the  line  u  v  will  depend 
on  the  position  of  rivets  in  the  head  of  the  boiler,  and  this  distance  should  be  so 
adjusted  that  the  stud  or  studs  whii-li  fasten  the  quadrant  B  to  the  boiler  will  be  clear 
of  the  rivet  heads. 

366.  The  following  figures  represent  a  throttle-lever  arrangement  for  engines,  in 
which  the  throttle  rod  passes  through  the  side  of  the  dome.  Similar  letters  in  the 
different  views  indicate  the  same  details. 


360 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


•  HI 

,1 

5\ 

ft 

t 

*s 
\ 

<-vW 

•7* 

k 
i-< 

} 

A 

ffl 


Fig.  525  represents  this  throttle-lever  arrangement  as  seen  from  the  back  end  of 
the  boiler.  Fig.  526  is  a  plan  of  the  same ;  Fig.  527,  a  section  of  the  steam-gauge 
stand,  stuffing-box,  gland,  and  steam-pipe  connecting  the  steam-gauge  stand  to  the 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


361 


dome ;  Fig.  528  shows  the  throttle  rod,  and  Fig.  529  the  link  used  for  connecting  the 
throttle  lever  to  the  steam-gauge  stand. 

The  steam-gauge  stand  marked  A  is  often  made  of  cast-iron,  sometimes  of  brass, 


tl'^,Threadi.fo  one. inch., 


Fig.  527 

and  is  bolted  to  the  top  of  the 
boiler  by  means  of  two  studs  pass- 
ing through  the  holes  a  a.  The 
throttle  rod  G  works  in  the  brass 
stuffing-box  B,  which  is  fastened 
to  the  steam-gauge  stand  by  means 
of  the  two  studs  b  b ;  these  studs 
are  also  used  for  tightening  the 
brass  gland  C.  The  stuffing-box 
is  bored  out  at  one  end  to  receive 
the  hemp  packing  and  the  gland ; 
the  other  end  of  the  stuffing-box 
is  tapped,  and  the  wrought-iron 
tube  or  pipe  K  screwed  into  it. 
This  pipe  is  about  ^  of  an  inch 
thick,  and  is  the  same  kind  of 
tubing  as  used  for  water  grates,  to  which  we  shall  refer  later  on.  The  other  end  of 
the  pipe  K  is  screwed  into  a  small  brass  flange ;  the  latter  is  riveted  to  the  outside 


362  MODERN  LOCOMOTIVE  CONSTRUCTION. 

of  the  dome,  and  is  completely  covered  by  the  dome  casing.  The  pipe  K  is  not  bored 
out ;  its  inner  diameter  is  somewhat  larger  than  the  diameter  of  throttle  rod  G,  so  as 
to  give  the  latter  ample  freedom  for  its  motion.  The  purpose  of  the  pipe  K  is  to  cover 
and  protect  the  throttle  rod,  and  in  the  meantime  bring  the  stuffing-box  B  and  gland 
C — which  are  required  in  any  case — within  easy  reach  of  the  engineer. 

In  this  design  the  jaw  H  and  the  jaw  at  the  opposite  end  of  the  throttle  rod  are 
keyed  to  the  latter,  as  indicated  in  Fig.  528 ;  this  way  of  fastening  the  jaws  to  the 
throttle  rod  differs  a  little  from  the  manner  of  fastening  similar  jaws  on  rods,  as  pre- 
viously illustrated.  The  wrought-iron  link  I  forms  a  connection  between  the  throttle 
lever  E  and  the  lug  C2  cast  to  the  steam-gauge  stand  A,  and  serves  as  a  fulcrum 
for  the  former. 

The  manner  of  locking  the  lever  E  differs  greatly  from  any  of  the  previous  de- 
signs. The  upper  side  of  the  jaw  H  (Figs.  525,  526)  is  extended  sideways  and  formed 
into  a  circular  rack ;  the  pitch  of  the  teeth  is  very  fine,  so  as  to  obtain  a  regulation  of 
the  lift  of  the  throttle  valve  as  close  as  possible  to  the  requirements  of  the  engine ;  yet 
with  this  arrangement  it  will  be  difficult,  if  not  impracticable,  to  obtain  as  close  a 
regulation  as  with  the  design  shown  in  Fig.  518.  A  steel  latch  L,  having  three  or  four 
teeth  cut  in  its  end,  engages  with  the  rack.  The  lug  forged  to  the  bottom  of  the 
latch,  and  sliding  in  a  slot  cut  through  the  lever,  serves  as  a  guide  for  the  latch ;  the 
link  F  connects  the  latch  to  the  handle  M,  which,  in  being  pressed  towards  the  throttle- 
lever  handle,  disengages  the  latch  from  the  rack,  and  leaves  the  lever  free  to  move. 
In  moving  the  lever  to  and  fro,  the  latch  L  will  move  faster  than  the  pin  d,  and  it  is 
on  account  of  the  difference  between  the  rates  of  these  motions  that  when  the  latch  L 
engages  with  the  circular  rack  the  lever  is  locked. 

The  steam-gauge  is  fastened  to  the  upper  part  of  the  stand  A ;  the  center  /of  this 
part  coincides  with  the  center  of  the  steam-gauge. 

The  handle  Z>  J9,  Figs.  525,  526,  is  for  the  purpose  of  opening  one  of  the  safety 
valves,  or  regulating  the  pressure  on  the  same ;  the  pin  li  connects  a  spring  balance — 
not  shown — to  the  lever  D ;  this  spring  balance  stands  in  a  vertical  position,  with  its 
upper  end  attached  to  the  safety-valve  lever;  the  handle  D  swings  on  the  pivot  i, 
which  is  cast  on  the  back  of  the  steam-gauge  stand  A ;  the  pawl  N  engages  with  the 
teeth  cut  on  the  edge  of  the  steam-gauge  stand,  and  prevents  the  lever  D  from  moving 
upwards.  In  pulling  the  lever  downwards,  the  pressure  on  the  safety  valve  will  be 
increased.  The  safety  valves  and  spring  balance  will  be  described  later  on ;  all  that 
we  need  to  say  here  is  that  two  safety  valves  are  always  used  for  a  locomotive  boiler, 
and  it  is  only  one  of  these  that  can  be  released,  or  the  pressure  upon  it  changed  by  the 
lever  D ;  the  other  safety  valve  is  or  should  be  beyond  the  control  of  the  engineer. 

Fig.  530  represents  a  wrought-iron  steam-gauge  lamp  bracket  for  the  throttle 
valve  gear  shown  in  Fig.  525 ;  its  end  p  is  inserted  in  the  hole  g  in  the  lower  part  of 
the  stand  A,  and  fastened  there ;  the  steam-gauge  lamp  is  screwed  on  to  the  end  I  of 
the  lamp  bracket. 

367.  In  this  design  of  throttle-valve  gear  the  pitch  line  of  the  teeth  in  the  circular 
rack  must  be  described  from  the  center  d,  Fig.  526,  and  not  from  a  center  lying  to 
the  left  of  it — for  instance,  such  as  we  were  compelled  to  find  by  the  construction 
shown  in  Fig.  524. 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


363 


For  obtaining  the  length  of  the  circular  arc — that  is,  the  distance  from  h  to  i, 
Fig.  531 — we  simply  describe  from  the  center  d  the  arc  h  i,  which  contains  the  outer 
extremities  of  the  teeth,  and  then  proceed  as  follows : 

Let  the  point  d,  Fig.  531,  represent  the  position  of  the  center  of  the  pin  when  the 
throttle  valve  is  closed.  Through  the  point  d  draw  the  center  line  m  n  of  the  valve- 
rod,  and  on  this  line  mark  off  a  point  d% ;  the  distance  between  the  points  d  and  d2 
must  be  equal  to  the  distance  through  which  the  center  d  of  the  pin  will  travel  to  open 
the  throttle  valve  fully ;  this  distance  is  found  as  explained  in  Art.  365,  and  is  usually 
about  two  inches,  or  a  little  more ;  in  some  cases  it  may  be  as  much  as  2£  inches. 


Fig.  SSI 

Lay  off  a  point  c  to  represent  the  center  of  the  lug  cast  on  the  steam-gauge  stand ;  the 
location  of  the  point  c  must,  of  course,  correspond  relatively  to  the  position  of  the 
point  d.  From  the  point  c  as  a  center,  and  with  a  radius  equal  to  the  distance  between 
the  centers  of  the  holes  in  the  link  7  (Fig.  529),  describe  a  short  arc  r  s ;  also  from  the 
point  d  as  a  center,  and  with  a  radius  equal  to  the  distance  between  the  centers  d  and  e — 
that  is,  the  given  distance  between  the  centers  of  the  holes  in  the  lever  E — describe  an 
arc  t  u  cutting  r  s  in  the  point  e ;  the  straight  line  joining  the  points  c  and  e  will  be  the 
center  line  of  the  link  /.  Through  the  points  c  and  d  draw  a  straight  line,  and  prolong 
it  towards  L,  cutting  the  arc  h  i  in  the  point  #.  Now  the  center  line  e  L  indicates  one 
of  the  extreme  positions  of  the  throttle  lever  E  (that  is,  when  the  throttle  valve  is 


364  MODERN  LOCOMOTIVE  CONSTRUCTION. 

closed),  and  theoretically  the  arc  h  i  need  not  extend  beyond  the  edge  v  of  the  latch  L ; 
but,  in  order  to  allow  for  wear  and  inaccuracies  in  fitting,  the  distance  from  g  to  h  is 
usually  such  as  to  extend  $  of  an  inch  beyond  v ;  hence  the  total  distance  from  g  to  h 
measured  on  the  arc  h  i  will  be  about  If  inches. 

Through  the  point  d  draw  a  line  d  k  perpendicular  to  m  n,  cutting  the  arc  h  i  in  the 
point  q,  and  thus  obtain  the  distance  Ji  q — that  is,  the  distance  from  the  point  h  to  the 
line  d  k,  measured  on  the  arc  /*  i. 

If  now  a  line  w  x,  perpendicular  to  m  n,  be  drawn  through  the  point  w — that  is,  the 
middle  of  the  distance  d  d2 — and  the  line  w  x  passes  through  the  point  e,  as  shown,  then 
the  motion  of  the  lever  E  will  be  symmetrical  on  each  side  of  the  line  e  w ;  under  these 
conditions  we  have  only  to  make  i  q  equal  h  q,  and  thus  obtain  the  whole  length  of  the 
circular  rack  Ji  i. 

If  the  line  w  x  does  not  pass  through  the  point  e,  then,  the  point  h  having  been 
found  as  before,  the  point  i2  will  be  found  in  the  following  manner : 

Through  the  point  d2  draw  the  line  d2  k2,  perpendicular  tomn;  also  from  d2  as  a 
center,  and  with  a  radius  equal  to  d  e,  describe  an  arc  to  cut  the  arc  r  s ;  in  practice 
the  point  of  intersection  thus  found  will  generally  be  so  near  to  the  point  e  that  we 
may  consider  the  former  to  coincide  with  the  point  e.  Through  the  point  e  and  d2 
draw  a  straight  line,  and  prolong  it  towards  f2. 

From  the  point  d2  as  a  center,  and  with  a  radius  equal  to  d  g,  describe  the  arc  q2  i2, 
cutting  ef>  in  the  point  #,,  also  cutting  the  horizontal  line  d2  k2  in  the  point  q2.  Make 
g2  i2  equal  to  g  h,  add  the  arc  q2  i2  to  the  arc  h  q,  and  thus  obtain  the  whole  length  of 
the  circular  rack. 

368.  In  Art.  366  we  have  shown  a  throttle-valve  gear  attached  to  the  top  of  boiler, 
with  the  throttle  rod  passing  through  the  side  of  the  dome.  Figs.  532,  533,  534  repre- 
sent another  throttle- valve  gear,  also  attached  to  the  top  of  boiler,  but  differing  from 
the  former  in  having  the  throttle  rod  H  pass  through  the  top  of  the  boiler  instead  of 
passing  through  the  side  of  the  dome. 

Fig.  532  represents  the  relative  positions  of  the  steam-gauge  stand  A,  the  stuffing- 
box  B  with  gland  C,  and  the  throttle  lever  E.  Fig.  533  simply  represents  a  plan  of 
the  stuffing-box  and  gland,  the  throttle  lever,  and  the  notched  quadrant  D ;  the  steam- 
gauge  stand  is  not  shown  in  this  figure. 

Fig.  534  represents  the  relative  positions  longitudinally  of  the  stuffing-box  and 
steam-gauge  stand ;  and  Fig.  535  represents  a  plan  of  the  latter. 

The  throttle  rod  If  stands  in  a  vertical  position.  The  ends  of  the  throttle  rod 
which  pass  through  the  lever  E  and  the  crank  I  are  cut  square. 

In  Fig.  532  the  line  d  is  the  center  line  of  the  boiler,  and  since  the  center  line  of 
the  throttle  pipe  in  the  dome  coincides  with  the  line  d,  and  since  the  center /of  the 
pin  through  the  small  crank  I  should  also  coincide,  or  nearly  so,  with  d,  it  follows  that 
the  stuffing-box  must  be  placed  on  one  side  the  center  line  d. 

The  joint  between  the  stuffing-box  and  top  of  boiler  is  a  ground  ball  joint ;  around 
it,  the  thickness  of  metal  is  increased  by  riveting  a  small  plate  K  to  the  inside  of  the 
boiler,  thereby  obtaining  a  sufficient  depth  of  metal  for  the  threads  on  the  studs  o,  a2,  a2, 
which  fasten  the  stuffing-box  to  the  boiler ;  the  studs  a2  a-i  are  made  long  enough  to 
take  hold  of  the  gland  C. 


-^ a 


366  MODERN  LOCOMOTIVE   CONSTRUCTION. 

The  steam-gauge  stand  A  is  fastened  to  the  top  of  boiler  with  two  studs,  b  b.  The 
depth  e  of  the  steam-gauge  stand  flange  may  seem  to  be  excessive ;  this  depth,  as  well 
as  the  distance  from  the  top  of  the  stuffing-box  flange  to  the  top  of  boiler,  will  be  gov- 
erned by  the  following  conditions : 

The.  designs  of  throttle- valve  gears  shown  in  Figs.  525  and  532  are  generally 
required  for  boilers  which  extend  nearly  to  the  rear  end  of  the  cab ;  the  portion  of  the 
boiler  inside  of  cab  must  be  lagged,  as  well  as  the  outer  portion,  so  as  to  prevent  a  loss 
of  heat  by  radiation,  and  also  to  prevent  the  cab  from  becoming  uncomfortably  hot  for 
the  engineer.  The  distance  between  the  outer  face  of  lagging  and  the  boiler  usually 
varies  from  about  Ij  to  2  inches ;  now,  in  order  to  make  a  nice  and  easy  finish  of  the 
lagging  around  the  steam-gauge  stand  and  stuffing-box,  the  upper  faces  of  their  flanges 
are  made  to  extend  about  \  inch  beyond  the  lagging,  hence  the  excessive  depth  e  of 
the  steam-gauge  stand  flange  and  the  seemingly  unnecessary  height  of  the  stuffing- 
box  flange. 

The  quadrant  D  is  bolted  to  the  steam-gauge  stand ;  this  quadrant  must,  of  course, 
be  described  from  the  center  of  the  throttle  rod  H.  It  will  be  noticed  that  in  this  design 
the  quadrant  D  is  made  to  pass  through  the  throttle  lever  E,  instead  of  being  placed 
below  or  above  it,  as  shown  in  some  of  the  other  designs  of  throttle-valve  gear.  The 
manner  of  locking  the  throttle  lever  E  is  so  plainly  shown  in  Fig.  533  that  an  explana- 
tion is  unnecessary. 

The  arc  y  h  i,  Fig.  533,  represents  the  path  of  the  end  of  the  throttle  lever ;  the 
distance  between  its  extremities  g  and  i  is  usually  about  12  inches,  and  should  not 
exceed  18  inches.  Now,  to  keep  the  movement  of  the  end  of  the  throttle  lever  within 
these  limits,  and  in  the  meantime  give  the  throttle  valve  the  required  lift,  no  more  and 
no  less,  we  must  assign  a  suitable  length  to  the  crank  I  in  Fig.  532.  But  in  many 
cases  the  boiler  braces  will  determine  the  position  of  the  stuffing-box  B,  which  will  also 
fix  the  length  of  the  crank  7;  the  length  thus  found  may  not  be  suitable  for  keeping 
the  movement  of  the  throttle  lever  within  the  given  limits ;  under  these  conditions  we 
must  give  such  lengths  to  the  arms  of  bell  crank  B,  Fig.  493,  as  will  produce  the 
desired  results.  The  given  limit  of  the  movement  of  the  throttle  lever  will  also,  in 
many  cases,  determine  the  distance  between  the  holes  e  and  d  in  Fig.  531. 

369.  We  have  already  seen  that  the  design  of  the  throttle- valve  gear  shown  in 
Figs.  525,  526  is  used  on  engines  whose  boilers  extend  nearly  to  the  rear  end  of  the 
cab.  This  class  of  engines,  or  the  class  of  engines  in  which  the  throttle-rod  passes 
through  the  side  of  dome,  generally  present  first-class  opportunities  for  making  pro- 
visions for  attaching  the  various  kinds  of  valves  and  cocks  without  screwing  each  one 
directly  into  the  boiler  shell.  Consequently,  in  many  of  the  locomotives  of  these  classes 
built  in  recent  years,  we  find  a  throttle- valve  gear  like  that  shown  in  Figs.  536,  537,  or 
others,  very  similar  in  design  to  the  one  here  shown.  In  these  illustrations  we  have 
only  represented  the  most  prominent  features  of  this  design  of  throttle- valve  gear ;  for 
the  sake  of  simplicity  the  throttle  lever,  with  its  attachments  for  locking  it,  is  not 
shown ;  in  fact,  any  one  of  the  throttle  levers  previously  illustrated,  with  only  a  slight 
modification  in  a  few  of  them,  can  be  used  in  this  kind  of  gear.  Its  principal  feature 
is  the  steam  stand  B,  which  is  simply  a  rectangular  box,  generally  made  of  brass. 
In  Fig.  536  we  see  a  longitudinal  section  of  this  box,  and  Fig.  540  shows  a  cross- 


/HW3~T         „  " 
/    ~pj 

:SJ£ 

B'  PS" 


368  MODERN  LOCOMOTIVE   CONSTRUCTION. 

section  of  the  same.  These  figures  plainly  indicate  that  the  steam  stand  B  is  divided 
into  two  compartments  or  chambers,  D  and  C;  the  chamber  C  nearly  surrounds  the 
chamber  D.  Communication  between  these  two  chambers  is  either  opened  or  closed  by 
means  of  the  valve  e. 

The  throttle  rod  H  passes  through  the  chamber  D,  and  also  through  the  heavy 
wrought-iron  pipe_EJ;  the  latter  forms  a  connection  between  the  steam  stand  B  and  the 
dome  G.  The  rear  end  of  the  steam  stand  is  bored  out  so  as  to  form  a  stuffing-box ;  h  is 
the  stuffing-box  gland.  The  end  i  of  the  throttle  rod  H  is  fastened  to  the  throttle 
jaw  (not  shown  here),  and  this  throttle  jaw  connects  to  the  throttle  lever. 

The  steam-pipe  E  is  connected  to  the  dome  by  means  of  a  thimble  K,  which  takes 
the  place  of  a  brass  flange.  A  part  of  the  outer  portion  of  this  thimble  is  hexagonal  in 
form,  the  remaining  outer  portion  is  threaded  and  screwed  into  the  dome  sheet.  A 
portion  of  this  thimble  is  tapped  and  receives  the  threaded  end  of  the  pipe  E.  This  is 
a  favorite  way,  in  locomotive  practice,  of  connecting  a  pipe  of  the  kind  here  shown  to 
the  boiler.  On  the  other  end  of  the  pipe  E  a  wrought-iron  sleeve  /  is  screwed.  A 
separate  view  of  this  sleeve  is  shown  in  Fig.  543,  and,  as  will  be  seen,  it  forms  a  ball 
joint  with  the  steam  stand ;  this  sleeve  is  held  against  the  stand  by  the  flange  F,  of 
which  a  separate  view  is  shown  in  Fig.  542. 

Fig.  537  represents  a  plan  of  the  steam  stand  B ;  here  it  is  plainly  seen  that  its 
sides  have  a  number  of  tapped  holes;  into  these  the  various  valves  and  cocks  are 
screwed,  which  otherwise  would  have  to  be  screwed  into  the  shell  of  the  boiler.  For 
instance,  the  holes  a  a  receive  the  injector  valves ;  b  b  receive  the  cylinder  oil-cups ; 
c  receives  the  blower  valve ;  and  s  the  steam-gauge  cock ;  the  hole  d  on  top  of  the  stand 
takes  the  brake  valve. 

Steam  is  conveyed  from  the  dome  through  the  pipe  E,  and  enters  the  chamber  D 
in  the  steam  stand ;  when  the  valve  e  is  open  the  steam  enters  the  chamber  C,  and 
supplies  all  the  valves  attached  to  the  steam  stand. 

Fig.  541  shows  the  spindle  of  the  valve  e  on  a  larger  scale ;  Fig.  538  shows  the 
front  end  of  the  steam  stand  B,  and  a  portion  of  the  steam-gauge  stand  A ;  Fig.  539 
shows  the  rear  end  of  the  stand  B,  with  stuffing-box  gland  h;  the  tapped  hole  g 
receives  the  fulcrum  for  the  throttle  lever.  In  Fig.  536,  8  is  the  steam-gauge,  and  L 
the  steam-gauge  lamp. 

We  believe  this  arrangement  to  be  one  of  the  best  and  neatest  in  use.  All  the 
holes  in  the  steam  stand  B  can,  of  course,  be  drilled  in  a  machine,  and  all  the  valves, 
etc.,  fitted  in  it  before  it  is  taken  into  the  erecting  shop,  and  therefore  less  time  and 
labor  will  be  required  for  attaching  the  different  valves,  cocks,  etc.,  to  the  engine  than 
must  be  expended  when  each  one  of  the  valves  has  to  be  fitted  directly  into  the  boiler 
shell.  But  besides  this  advantage,  the  steam  stand  possesses  another  one,  namely,  with 
the  valve  e,  the  steam  can  be  at  once  shut  off  from  all  the  valves  attached  to  the  stand, 
consequently,  if  one  of  these  valves  gets  out  of  order,  it  can  be  repaired  with  full  steam 
pressure  in  the  boiler. 

The  relative  position  of  the  steam  stand  B  will  greatly  depend  upon  the  position 
of  the  reverse  lever ;  it  must  be  placed  in  a  position  which  will  bring  the  throttle-valve 
lever,  as  well  as  the  reverse  lever,  within  easy  reach  of  the  engineer ;  consequently  we 
often  find  that  this  design  of  throttle-valve  gear  is  placed  quite  a  distance  in  front  of 


MODERN  LOCOMOTIVE   COXSTKVCTION. 


369 


the  back  head  of  the  boiler,  and  differs  from  the  position  generally  assigned  to  the  throttle- 
valve  gear  in  ordinary  eight-wheeled  passenger  engines.  In  fact,  in  the  latter  class  of 
engines  the  boiler  seldom  extends  more  than  12  to  15  inches  into  the  cab ;  hence,  the 
steam  stand  illustrated  in  Figs.  536,  537  is  not  suitable  for  this  class  of  engines,  because 
there  is  no  room  for  it.  The  short  extension  of  the  boiler  into  the  cab,  and  general 
design  of  passenger  engines,  necessitate  the  use  of  throttle-valve  gears  such  as  have 
been  illustrated  in  Figs.  514,  517,  519.  When  one  of  the  latter  class  of  throttle-valve 
gears  has  to  be  used,  the  steam-gauge  stand  becomes  an  independent  fixture,  and,  for 
the  sake  of  convenience  to  the  engineer,  it  is  generally  fastened  to  the  curved  part  of 
the  back  head  of  the  boiler. 

STEAM-GAUGE  STAND. 

370.  Fig.  544  represents  a  steam-gauge  stand  for  passenger  engines  and  for  that 
class  of  engines  whose  boiler  projects  but  little  into  the  cab.  This  stand  is  arranged 
for  a  steam-gauge  and  clock,  the  latter 
being  placed  above  the  former.  In  this 
figure  we  also  see  an  outside  view  of  the 
spring  balance  S  connected  to  the  lever 
C.  The  rod  E  extends  upwards  and  is 
connected  to  the  safety-valve  lever.  A 
portion  of  the  lever  C  is  represented 
in  section,  so  as  to  show  the  spiral 
spring  underneath  the  pawl  D.  The 
spiral  spring  keeps  the  pawl  engaged 
with  the  circular  rack  e,  and  prevents 
the  lever  C  from  being  pulled  upwards 
by  the  tension  of  the  spring  balance. 
The  lever  C  swings  on  the  pivot  f,  which 
is  cast  to  the  steam-gauge  stand; 
hence,  in  pulling  the  lever  C  down- 
wards the  tension  of  the  spring  will  be 
increased,  and  therefore  the  force  which 
presses  the  safety  valve  against  the 
seat  will  also  be  increased.  Pressing  the 
pawl  D  towards  the  lever  C  disengages 
the  pawl  and  rack,  allowing  the  lever  to 
move  upwards,  thus  enabling  the  engi- 
neer to  blow  off  steam  when  necessary. 

Another  view  of  the  pawl  is  shown 
at  D2.  The  nuts  g  g  fasten  the  steam- 
gauge  lamp  to  the  stand.  Fig.  545  rep- 
resents a  plan  of  the  lever  (7,  and  a 

JT  w/»  5-/-A 

section  of  the  steam-gauge  stand. 

Fig.  546  represents  a  side  view  of  the  same  steam-gauge  stand.     This  side  view  is 
drawn  to  a  larger  scale,  so  as  to  enable  us  to  illustrate  more  distinctly  a  section  of  the 


370 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


spring  balance.     This  spring  balance  consists  of  an  outer  and  inner  casing,  a  spring 
which  connects  the  two,  and  nuts  for  regulating  the  tension  of  the  spring.     The 


Fig.  550 


Fig.  54G 


__;<±2 


?V 


3*=-~ 
^% 


Fig.  5.48  M 


L- .~ 


A 
Fig.  547 


casings  are  made  of  brass  tubes.  The  outer  casing  is  open  at  the  bottom,  and  a  small 
brass  head  B  is  brazed  in  its  upper  end ;  this  head  is  bored  out  to  receive  the  brass 
nut  H,  of  which  separate  views  are  shown  in  Fig.  549.  The  inner  casing  is  open  at 


MODERN  LOCOMOTIVE  CONSTRUCTION.  371 

the  top,  and  closed  at  the  bottom  by  the  brass  head  A  brazed  to  it.  A  separate  view  of 
the  inner  casing  is  shown  in  Fig.  547 ;  a  portion  of  the  upper  end  of  this  casing  is  cut 
off  for  convenience  in  putting  the  spring  balance  together.  Fig.  548  represents  the 
spiral  springs,  one  placed  inside  of  the  other,  and  fastened  to  the  brass  end-pieces  L 
and  M;  a  plan  of  the  end-piece  L  is  shown  at  L^  and  a  plan  of  the  lower  end-piece 
M  is  shown  at  M2 ;  in  this  figure  the  ends  of  the  two  springs  are  also  seen.  The  piece 
L  is  fastened  to  the  head  of  the  outer  casing  by  means  of  two  screws,  as  shown  at  $2, 
Fig.  546,  which  is  another  section  of  the  upper  part  of  the  outer  casing ;  the  end-piece 
M  is  attached  to  the  head  at  the  bottom  of  the  inner  casing. 

The  upper  end  of  the  outer  casing  fits  in  the  recess  of  the  cap  7),  of  which  separate 
views  are  shown  in  Fig.  550.  This  cap  is  not  fastened  to  the  casing.  The  feather  p, 
shown  in  Fig.  550,  engages  with  the  groove  cut  in  the  outer  surface  of  the  nut  H; 
hence,  in  turning  the  cap,  the  nut  must  also  turn,  causing  an  increase  or  decrease  in 
the  tension  of  the  spring.  The  nut  F,  Fig.  546,  is  simply  a  jam  nut. 

Examining  the  section  in  Fig.  546,  it  will  be  seen  that  the  head  of  the  nut  H  bears 
against  the  under  side  of  the  upper  end-piece  L  of  the  springs ;  consequently  the  outer 
casing  will  not  be  subjected  to  any  vertical  stress  due  to  the  pressure  on  the  safety 
valve ;  the  two  small  screws  connecting  the  head  of  the  outer  casing  and  the  end-piece 
L  are  simply  for  the  purpose  of  preventing  the  piece  L  from  turning,  thereby  keeping 
the  springs  free  from  any  torsional  strain. 

The  tension  of  the  springs  can,  of  course,  be  regulated  to  a  limited  extent  by 
means  of  the  lever  (7,  of  which  other  views  are  shown  in  Figs.  544,  545.  The  nut  II 
and  cap  D  are  for  the  purpose  of  closer  adjustment  of  the  tension. 

In  Fig.  546,  0  represents  the  clock,  G  the  steam-gauge,  and  L  the  lamp. 

SAFETY  VALVE,  DOME,   AND  CASING. 

371.  Fig.  551  represents  a  common  safety  valve,  its  attachments,  and  a  portion  of 
the  dome  top  marked  D.  Occasionally  the  opening  for  the  common  safety  valve 
in  dome  top  is  bushed  with  brass,  but  generally  it  is  not  bushed.  The  safety  valve 
A  is  made  of  brass;  it  consists  of  a  hollow  cone  with  four  wings,  a  a,  cast  to  it, 
which  guide  the  valve  in  the  opening.  Frequently  the  valve  seat  is  made  flat,  leaving 
only  a  bearing  of  -fa  inch  all  around,  as  shown  in  the  illustration,  but  this  is  not  the 
best  form.  The  seat  should  have  an  inclination  of  45  degrees  to  the  center  line  of 
its  axis,  thereby  obtaining  an  additional  face-to-face  metal  impingement,  which  insures 
tightness  under  a  high  boiler  pressure.  This  is  an  important  matter,  particularly 
for  the  higher  pressures  as  are  now  used  in  locomotives,  because  with  an  increased 
metal  impingement  the  valve  will  keep  tight  to  a  limit  nearer  to  the  blowing-off  point 
than  those  with  flat  seats.  An  angle  less  than  45  degrees  would  be  still  better  to 
insure  against  leakage,  but  with  this  comes  the  danger  of  the  valve  sticking  to  its  seat. 
Hence,  the  seat  beveled  to  an  angle  of  45  degrees  we  believe  to  be  the  best. 

The  surface  of  the  seat  should  not  be  conical ;  it  should  be  spherical,  so  that  the 
valve  will  always  be  tight  even  when  there  is  not  the  proper  alignment  of  motion  from 
tlif  want  of  accuracy,  of  workmanship,  or  from  wear.  The  radius  of  this  surface  is 
found  in  the  following  manner :  Draw  the  valve  seat  as  shown  in  Fig.  551A ;  bisect 


372 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


MODERN  LOCOMOTIVE  CONSTRUCTION.  373 

a  b  and  c  d  by  perpendiculars,  cutting  each  other  in  the  point  o;  then  o  a  will  be 
the  required  radius. 

Two  views  of  the  safety-valve  lever  F  are  shown  in  Fig.  551,  F2  being  the  plan. 
The  rounded  end  of  the  wrought-iron  valve  spindle  B  sets  in  the  hollow  part  of  the 
valve.  The  wrought-iron  fulcrum  C  is  screwed  into  the  dome  top.  Another  view 
of  this  fulcrum  is  shown  at  (72.  The  rod  E  is  a  portion  of  the  rod  marked  E  in 
Fig.  546. 

The  opening  in  the  top  of  dome  for  this  kind  of  safety  valve  is  three  inches  in 
diameter  for  all  locomotives,  excepting  very  small  ones,  say  with  cylinders  nine  or  ten 
inches  in  diameter ;  and  even  for  these  engines  the  safety-valve  opening  is  sometimes 
three  inches  in  diameter. 

This  form  of  safety  valve  is  often  adopted  when  the  dome  is  close  to  the  cab,  as 
shown  in  Fig.  552 ;  it  is  placed  on  the  right-hand  side  of  the  engine,  a  pop  being  placed 
on  the  other  side. 

When  the  dome  is  placed  on  the  center  of  the  boiler,  the  common  safety  valve 
here  shown  is  not  suitable,  because  its  lever  will  be  too  long,  and  therefore  a  pop  valve 
is  used  in  its  place.  Some  master-mechanics  use  the  latter  valve  exclusively  on  all 
engines.  The  advantage  claimed  for  the  common  safety  valve  with  the  spring  balance 
arranged  as  illustrated  is  that  it  can  be  adjusted  very  conveniently — without  going 
outside  of  the  cab — to  blow  off  at  any  desired  steam  pressure.  On  the  other  hand,  the 
pop  valve  has  a  greater  venting  capacity,  and  is  therefore  sometimes  preferred.  These 
valves  will  be  described  later. 

372.  Fig.  552  represents  a  section  of  a  portion  of  the  boiler  B  and  the  dome  D ;  it 
shows  plainly  the  position  of  the  steam-gauge  stand  in  the  cab,  also  the  manner  of 
connecting  the  spring  balance  S,  by  means  of  the  rod  E,  to  the  safety-valve  lever  F. 
The  whistle  lever  is  marked  G ;  details  of  the  whistle  and  connections  will  be  shown 
later.  We  also  see  the  position  of  the  throttle  pipe  in  the  dome,  and  the  general 
arrangement  of  the  throttle  gear.  The  throttle  lever,  which  is  not  shown,  connects 
to  the  jaw  7,  and  the  manner  of  opening  the  throttle  valve,  by  pulling  out  the  throttle 
rod  //,  can  readily  be  traced. 

The  steam-pipe  T  is  allowed  to  rest  on  the  crown  bars  V,  being  secured  in  posi- 
tion by  the  clamp  W,  which  is  bolted  to  the  side  of  the  dome.  We  also  see  the  rela- 
tive position  of  the  reverse  lever  />,  which  is  shown  in  full  gear  forward.  The  reverse 
lever  is,  with  veiy  few  exceptions,  always  placed  on  the  right-hand  side  of  the  engine ; 
it  is  shown  here  for  the  sake  of  completeness ;  had  we  strictly  followed  the  rules  of 
drawing  we  could  not  have  shown  the  reverse  lever,  because  that  side  of  the  engine 
is  cut  off. 

The  section  here  shown  is  that  of  a  switching  engine  with  a  saddle  tank  A  A,  but 
the  relative  positions  of  steam-gauge  stand,  reverse  lever,  and  throttle  lever  do  not 
differ  from  those  in  ordinary  eight-wheeled  passenger  engines.  In  the  latter  class  of 
engine  we  have,  of  course,  no  tank  on  the  top  of  boiler,  and  therefore  the  dome  casing 
extends  to  the  top  of  the  lagging  K.  The  casing  here  shown  consists  of  a  cast-iron 
ring  M,  which,  in  passenger  engines,  is  fitted  to  the  top  of  the  lagging,  and  the  sheet- 
iron  or  brass  casing,  which  is  made  in  three  parts,  namely,  the  lower  ring  N,  the  body 
P,  and  the  upper  ring  0.  The  casing  is  not  fastened  to  the  dome,  but  is  kept  in  place 


374 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


by  the  lagging  around  the  dome,  to  which  it  is  fitted  pretty  closely,  and  can  readily  be 
lifted  off  when  repairs  to  the  boiler  become  necessary. 

In  our  illustration  the  boiler  extends  a  little  further  into  the  cab  than  it  generally 
does  in  passenger  engines.  In  some  of  these  engines  the  lagging  around  the  boiler 
extends  only  up  to  the  cab ;  in  others  it  extends  to  the  back  head  of  the  boiler. 


TENSION   ON   SAFETY-VALVE   SPBINGS. 

373.  The  tension  on  the  springs  in  the  spring  balance  S  (Fig.  546),  the  length  of 
safety-valve  lever,  etc.,  are  determined  by  well-known  rules.  Before  we  give  these 
rules,  let  us  first  establish  a  formula  from  which  they  may  be  derived ;  such  a  course 
will  give  us  a  clearer  conception  of  them. 

In  Fig.  553  the  line  e  e2  represents  the  center  line  of  the  rod  marked  E  in  Figs. 
546, 551.  This  rod  is,  of  course,  subjected  to  a  tensile  stress  due  to  the  steam  pressure 

on  the  safety  valve,  and  this  stress  is 
transmitted  to  the  spring  balance; 
hence  we  may  say  that  the  tension 
on  the  spring  is  equal  to  the  stress 
on  the  rod  whose  center  line  is  rep- 
resented by  the  line  e  ez  in  Fig.  553. 
The  weight  of  the  safety-valve  lever, 
weight  of  valve,  and  weight  of  valve 
spindle  will,  of  course,  help  to  resist 
the  steam  pressure  on  the  valve,  and  will  therefore  re- 
duce the  tension  on  the  springs,  and  in  some  cases  this 
reduction  in  the  tension  will  be  considerable.     For  the 
sake  of  simplicity  we  shall,  in  the  first  place,  give  the 
rules  in  which  the  weights  of  the  safety-valve  lever, 
valve,  and  valve  spindle  are  neglected. 

Let  L  represent  the  distance  in  inches  from  the  center  of  the  fulcrum  to  the  center 
line  e  e2 ;  this  distance  is  often  called  the  length  of  the  safety-valve  lever. 

B,  the  distance  in  inches  from  the  center  of  the  fulcrum  to  the  center  of  the  valve 
spindle. 

T,  the  tension  in  pounds  on  the  springs. 

A,  the  area  in  square  inches  of  the  safety  valve ;  this  area  must  always  be  taken 
equal  to  the  cross-sectional  area  of  the  safety-valve  opening  D. 

P,  the  pressure  of  the  steam  in  pounds  per  square  inch  of  the  safety-valve  area. 

W2,  the  weight  in  pounds  of  the  safety  valve  and  its  spindle. 

W3J  the  weight  in  pounds  of  the  safety-valve  lever. 

C,  the  distance  in  inches  from  the  center  of  fulcrum  to  the  center  of  gravity  G  of 
the  lever. 

In  neglecting  the  weight  of  the  safety-valve  lever,  valve,  and  spindle,  the  symbols 
Wft  W3,  and  C  will  not  be  used ;  we  have  given  them  here  so  as  to  make  our  table  of 
symbols  complete. 

The  total  steam  pressure  in  pounds  on  the  safety  valve  is  evidently  equal  to  the 


Fig.  5SS 


MODERN  LOCOMOTIVE  CONSTRUCTION.  375 

product  obtained  by  multiplying  the  steam  pressure  P  per  square  inch  by  the  area  A 
of  the  valve ;  hence  the  total  pressure  on  the  valve  is  equal  to  P  x  A.  But  this  total 
steam  pressure  acts  with  a  leverage  B,  and  is  resisted  by  the  tension  T  acting  with  a 
leverage  L.  Now  in  order  to  compare  the  intensity  with  which  the  pressure  on  the 
valve  acts,  with  the  intensity  with  which  the  tension  acts,  we  must  find  the  moment  of 
each  of  these  forces. 

The  moment  of  the  total  steam  pressure  acting  on  the  valve  is  equal  to  P  x  A  x  B. 

The  moment  of  the  tension  is  equal  to  T  x  L. 

When  the  steam  pressure  on  the  valve  is  just  sufficient  to  raise  the  valve,  we  have 
then  a  condition  of  equilibrium,  in  which  the  moment  of  the  total  steam  pressure  is 
equal  to  the  moment  of  the  tension,  and  these  conditions  are  represented  by  the 

formula 

PxAxB=TxL. 

Putting  this  formula  in  words,  we  have :  The  product  obtained  by  multiplying 
the  steam  pressure  per  square  inch  by  the  area  in  square  inches,  and  by  the  distance 
from  the  center  of  the  valve  to  the  center  of  the  fulcrum,  is  equal  to  the  product 
obtained  by  multiplying  the  tension  in  pounds  by  distance  from  the  line  of  action  of 
the  tension  to  the  center  of  the  fulcrum. 

From  this  formula  we  derive  the  well-known  rules  which  will  enable  us  to  solve 
any  problem  relating  to  the  safety  valve,  and  these  rules  hold  true  when  a  weight  is 
substituted  for  the  springs,  as  in  safety  valve  for  stationary  boilei's ;  all  we  need  to 
remember  is  that  the  center  of  the  weight  will  lie  in  the  line  e  e2.  We  must  also  bear 
in  mind  that,  in  the  following  rules,  the  weight  of  the  safety-valve  lever,  valve,  and 
spindle  are  neglected. 

EXAMPLE  116. — The  length  L  (Fig.  553)  is  38 £  inches ;  the  distance  B  from  the 
center  of  the  valve  to  the  center  of  the  fulcrum  is  3  J  inches ;  the  steam  pressure  P 
per  square  inch  is  120  pounds ;  the  safety  valve  is  3  inches  diameter.  What  will  be 
the  tension  on  the  springs  ? 

The  area  A  of  a  safety  valve  3  inches  diameter  is  7.06  square  inches.  Here,  then, 
we  have  given  P,  A,  B,  and  L ;  it  is  required  to  find  T. 

We  know  that 

PxAxB=TxL; 
hence,  to  find  T,  we  have 

PxAxB _ 

L  ••  T,  (a) 

which  reads : 

RULE  87. — The  steam  pressure  in  pounds  per  square  inch,  multiplied  by  the  area 
in  square  inches  of  the  safety  valve,  and  this  product  again  multiplied  by  the  distance 
in  inches  from  the  center  of  the  valve  to  the  center  of  the  fulcrum,  and  the  last  prod- 
uct divided  by  the  length  of  the  lever  in  inches,  will  give  the  tension  in  pounds  of  the 
springs. 

Substituting  the  numerical  values  for  the  symbols  in  formula  (a),  we  have 

120  x  7.06  x  3.5 

,,s  -         -  =  <7  pounds 

for  the  tension  of  the  springs. 

EXAMPLE  117.— The  tension  T  is  77  pounds;  the  distance  B  (Fig.  553)  is  3£  inches; 


376  MODERN  LOCOMOTIVE  CONSTRUCTION. 

the  area  A  of  the  valve,  7.06  square  inches  ;  steam  pressure  P,  120  pounds  per  square 
inch.    Find  the  length  L  of  the  safety-valve  lever. 

Here  we  have 

P  x  A  x  B 

-jr-    -  =  L,  (I) 

which  reads  : 

RULE  88.  —  The  steam  pressure  in  pounds  per  square  inch,  multiplied  by  the  area 
in  square  inches  of  the  safety  valve,  and  this  product  again  multiplied  by  the  distance 
from  the  center  of  the  valve  to  the  center  of  the  fulcrum,  and  the  last  product  divided 
by  the  tension  in  pounds,  will  give  a  quotient  which  is  numerically  equal  to  the  length 
of  the  lever  in  inches. 

Substituting  the  numerical  values  for  the  symbols  in  formula  (6),  we  have 

120  x  7.06  x  3.5 

—  ^—        -  =  38.5  inches 

for  the  length  of  the  safety-valve  lever. 

EXAMPLE  118.  —  The  tension  T  is  77  pounds  ;  length  L  of  the  safety-valve  lever, 
38  J  inches;  area  A  of  the  valve,  7.06  square  inches;  distance  B  from  the  center  of  the 
valve  to  the  center  of  the  fulcrum,  3j  inches.  Find  the  steam  pressure  per  square  inch 
on  the  safety  valve. 

Here  we  have  T  x  L 

A  x  B  =     '  ^ 

which  reads  : 

RULE  89.  —  The  product  obtained  by  multiplying  the  tension  in  pounds  by  the 
length  of  the  lever  in  inches,  divided  by  the  product  obtained  by  multiplying  the  area 
in  square  inches  of  the  valve  by  the  distance  in  inches  from  the  center  of  the  valve  to 
the  center  of  the  fulcrum,  will  give  a  quotient  which  is  numerically  equal  to  the  steam 
pressure  per  square  inch. 

Substituting  the  numerical  values  for  the  symbols  in  fonnula  (c),  we  have 

77  x  38.5 

7.06  x  3.5  =  119'97  P°unds 

steam  pressure  per  square  inch  of  the  safety  valve. 

EXAMPLE  119.  —  The  tension  T  is  77  pounds  ;  length  of  lever,  38£  inches  ;  steam 
pressure,  120  pounds  ;  distance  from  the  center  of  the  valve  to  the  center  of  fulcrum, 
3£  inches.  Find  the  area  of  the  valve. 

Here  we  have 


which  reads  : 

RULE  90.  —  The  product  obtained  by  multiplying  the  tension  in  pounds  by  the 
length  of  the  safety-valve  lever  in  inches,  divided  by  the  product  obtained  by  multi- 
plying the  steam  pressure  per  square  inch  by  the  distance  from  the  center  of  the 
valve  to  the  center  of  the  fulcrum  in  inches,  will  give  a  quotient  which  is  numerically 
equal  to  the  number  of  square  inches  in  the  area  of  the  safety  valve. 

Substituting  the  numerical  values  for  the  symbols  in  formula  (d),  we  have 

77  x  38.5 

i  on  v  -3  r  =  7.0o  square  inches 

-L-jU    X    o.O 

in  the  area  of  the  valve. 


MODERN  LOCOMOTIVE   CONSTRUCTION.  377 

EXAMPLE  120. — The  tension  T  is  77  pounds ;  length  of  lever  L,  38J  inches ;  steam 
pressure  ]\  1'20  pounds ;  area  A,  7.06  square  inches.  Find  the  distance  from  the  center 
of  the  valve  to  the  center  of  the  fulcrum. 

Here  we«have 

T*L 

P*A-B> 

which  reads : 

RULE  91. — The  product  obtained  by  multiplying  the  tension  in  pounds  by  the 
length  of  the  lever  in  inches,  divided  by  the  product  obtained  by  multiplying  the 
steam  pressure  per  square  inch  by  the  area  in  square  inches  of  the  valve,  will  give  a 
quotient  which  is  numerically  equal  to  the  distance  in  inches  from  the  center  of  the 
valve  to  the  center  of  the  fulcrum. 

Substituting  the  numerical  values  for  the  symbols  in  formula  (c),  we  have 

77  x  38.5 
120  x  7.06  =  3'49  mches' 

which  is  the  distance  from  the  center  of  the  valve  to  the  center  of  the  fulcrum. 

When  the  weight  of  the  safety-valve,  lever,  etc.,  are  taken  into  account,  the 
formulas  become  a  little  more  complicated ;  but  if  the  foregoing  formulas  and  rules 
are  understood,  there  will  not  be  any  difficulty  in  forming  a  clear  conception  of  the 
rules  in  the  next  article. 

374.  In  the  following  rules  relating  to  safety-valve  problems,  the  weight  of  the 
valve,  lever,  and  spindle  is  to  be  taken  into  account.  These  weights  must  be  accu- 
rately ascertained,  either  by  actual  weighing  or  by  computation. 

The  weight  of  the  safety-valve  lever  will  act  on  the  valve  with  a  leverage  which 
is  equal  to  the  distance  from  the  center  of  gravity  of  the  lever  to  the  fulcrum.  In 
other  words,  we  assume  that  the  whole  weight  of  the  safety-valve  lever  is  concentrated 
at  its  center  of  gravity  G  (Fig.  553),  and  acts  with  a  leverage  C. 

When  the  safety-valve  lever  is  of  uniform  thickness  and  width  throughout, 
and  the  shape  of  one  end  exactly  like  that  of  the  opposite  end,  we  may,  for  all 
practical  purposes,  assume  the  center  of  gravity  to  lie  in  the  center  of  the  lever,  and, 
indeed,  it  will  be  there  exactly,  provided  no  holes  are  drilled  through  the  lever,  and  the 
metal  is  homogeneous. 

When  the  lever  is  not  of  uniform  thickness  throughout,  then,  in  order  to  find  the 
center  of  gravity,  the  lever  should  be  balanced  on  a  knife-edge,  and  when  in  equilib- 
rium, the  center  of  gravity  will  lie  in  a  vertical  line  drawn  across  the  lever  from  the 
knife-edge. 

When  the  lever  is  of  uniform  thickness  throughout,  but  not  of  uniform  width,  we 
may  also  find  the  center  of  gravity  by  balancing  the  lever  on  a  knife-edge.  This 
method  may  not  always  be  convenient;  in  such  cases  we  may  adopt  the  following 
method,  which  will  give  us  an  approximate  position  of  the  center  of  gravity ;  but  it 
should  be  distinctly  understood  that  this  method  is  only  applicable  to  levers  which 
have  a  uniform  thickness.  Nearly  all  safety-valve  levers  for  locomotives  are  of 
uniform  thickness,  with  the  exception  of  the  small  hub  around  the  center  line  <•  c.^ 
in  Fig.  553.  To  allow  for  this  hub,  assume  the  lower  edge  of  the  lever  to  extend 


378 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


lot 


clear  to  its  end  and  proceed  as  follows:  Cut  a  template  out  of  stiff  paper,  con- 
forming to  the  width  of  the  lever.  Anywhere  near  one  of  the  edges  of  this  template 
punch  a  small  hole,  for  instance  at  A  (Fig.  553a),  and  suspend  the  template  from  a  pin 
passed  through  this  hole,  allowing  the  template  to  have  plenty  of  freedom 
to  vibrate  ;  also,  from  the  same  pin  suspend  a  plummet-line,  and  along  it 
draw  a  pencil  line  A  B  on  the  template.  In  a  similar  way  suspend  the 
template  from  another  hole  C,  punched  anywhere  near  the  edge  opposite 
the  hole  A  ;  along  the  plummet-line  draw  on  the  template  another  pen- 
cil line  C  D,  cutting  A  B  in  the  point  G,  which  will  be  the  position 
of  the  center  of  gravity. 

The  preliminary  details  having  been  arranged,  we  are  in  a  position  to 
establish  a  formula,  from  which  all  subsequent  rules  relating  to  the 
safety-valve  can  be  deduced. 

Since  these  rules  will  be  interesting  examples  of  the  principle  of  mo- 
ments, it  may  be  advantageous  to  repeat  the  definition  of  the  moment  of 
a  force  given  in  Art.  256. 

The  moment  of  a  force  with  respect  to  a  point  is  the  product 
obtained  by  multiplying  the  force  by  the  perpendicular  distance  from  the 
point  to  the  direction  of  the  force. 

Now  the  point  referred  to  in  this  definition  is  the  center  of  the  ful- 
crum-pin in  Fig.  553.  About  this  point  there  are  four  forces  acting  — 
namely,  the  steam  pressure  acting  in  the  direction  of  the  vertical  line 
through  the  center  of  the  valve;  the  total  steam  pressure  is  evidently 
equal  to  P  x  A  (here  the  same  symbols  are  used  as  given  in  Art.  373), 
hence,  according  to  our  definition,  the  moment  of  this  pressure  about  the 
fulcrum  is  equal  to  P  x  A  x  B.  Second,  the  tension  acting  in  the  direction  of  the 
vertical  line  e  e2j  hence  its  moment  about  the  fulcrum  is  equal  to  T  x  L.  Third,  the 
weight  W2  of  the  valve  and  spindle  acting  in  a  direction  of  the  vertical  line  through 
the  center  of  the  valve,  hence  its  moment  about  the  fulcrum  is  equal  to  W2  x  B.  And 
lastly,  the  weight  W~3  of  the  lever,  acting  in  a  vertical  line  through  the  center  of 
gravity  G,  hence  its  moment  about  the  fulcrum  is  equal  to  W3  x  C. 

The  moment  P  x  A  x  B  acts  upwards,  all  the  other  moments  act  downwards  ;  and 
when  the  steam  pfessure  is  just  sufficient  to  raise  the  valve,  we  have  then  a  condition 
in  which  the  upward  moment  is  equal  to  the  sum  of  the  downward  moments.  Sum- 
ming up  the  downward  moments,  we  have  : 

Moment  of  the  tension  T  about  the  fulcrum  =  T  x  L. 

Moment  of  the  weight  Wz  about  the  fulcrum  =  TF2  x  B. 

Moment  of  the  weight  W3  about  the  fulcrum  =  W3  x  C. 

Sum  of  the  downward  moments  =  (T  x  L)  +  (W2  x  B)  +  (TF-,  x  C). 

Since  the  sum  of  these  moments  must  be  equal  to  the  upward  moment,  we  have  : 
P  x  A  x  B  =  (T  x  L)  +  (W2  x  B)  +  (W3  x  C).  (/) 


B 


Fig.  5S3a 


From  this  formula  all  rules  relating  to  the  safety  valve  can  be  deduced. 
EXAMPLE  121.  —  Tension  of  the  springs  T=  72.27  pounds. 


MODERN   LOCOMOTIVE    CONSTRUCTION  379 

Length  L  of  lever  =  38.5  inches. 

Weight  W3  of  lever  =  11  pounds. 

Distance  C  of  center  of  gravity  of  lever  from  the  fulcrum  =  15  inches. 

Weight  W2  of  valve  and  spindle  =  5  pounds. 

Distance  B  from  center  of  valve  to  fulcrum  =  3.5  inches. 

Area  A  of  the  valve  =  7.0G  square  inches. 

Find  the  steam  pressure  P  per  square  inch. 

Here  we  are  to  find  P  ;  hence  from  formula  (b)  we  obtain 


,        2  , 

~A^B~ 

which  reads  : 

RULE  92.  —  Add  the  moment  of  the  tension,  the  moment  of  weight  of  the  valve  and 
spindle,  and  the  moment  of  weight  of  the  lever;  divide  this  sum  by  the  product 
obtained  by  multiplying  the  area  of  the  valve  by  the  distance  from  the  fulcrum  to  the 
center  of  the  valve  ;  the  quotient  will  be  the  steam  pressure  per  square  inch  on  the 
valve. 

In  this  and  the  following  rules,  all  the  dimensions  should  be  taken  in  inches,  and 
the  weights  in  pounds. 

Substituting  the  numerical  values  for  the  symbols  in  formula  (g),  we  have 

(72.27  x  38.5)  +  (5  x  3.5)  +  (11  x  15) 
P  =  L  --  7.06X3*         -  '  =  119'9+  POUndS' 

EXAMPLE  122.  —  The  steam  pressure  per  square  inch  is  120  pounds  ;  the  weights 
and  dimensions  of  the  safety  valve  are  as  given  in  Example  121,  with  the  exception  of 
the  area  of  the  valve,  which  is  to  be  found. 

Here  we  have 


,  3  .,. 

PxB 

which  reads  : 

RULE  93.  —  Add  the  moment  of  the  tension,  the  moment  of  the  weight  of  valve  and 
spindle,  and  the  moment  of  the  lever  ;  divide  this  sum  by  the  product  obtained  by 
multiplying  the  steam  pressure  per  square  inch  by  the  distance  from  the  fulcrum  to 
the  center  of  the  valve  ;  the  quotient  will  be  the  area  of  the  valve  in  square  inches. 

Substituting  the  numerical  values  for  the  symbols  in  formula  (/*),  we  obtain 

(72.27  x  38.5)  +  (5  x  3.5)  +  (11  x  15) 

1  9Q  x  S  5  =         square  inches. 

EXAMPLE  123.—  The  steam  pressure  is  120  pounds,  the  weights  and  dimensions  of 
the  safety  valve  are  as  given  in  Example  121,  with  the  exception  of  the  distance  from 
the  fulcrum  to  the  valve,  which  is  to  be  found. 

Here  we  are  to  find  B,  hence 

G) 


2  3 

AxP 

which  reads  : 

RULE  94.  —  Add  the  moment  of  the  tension,  the  moment  of  the  weight  of  the  valve 
and  spindle,  and  the  moment  of  the  weight  of  the  lever;  divide  this  sum  by  the 


380  MODERN  LOCOMOTIVE   CONSTRUCTION. 

product  obtained  by  multiplying  the  area  of  the  valve  by  the  steam  pressure ;  the 
quotient  will  be  the  distance  from  the  fulcrum  to  the  center  of  the  valve  in  inches. 
Substituting  the  numerical  values  for  the  symbols  in  formula  (i),  we  obtain 

_  (72.27  x  38.5)  4-  (5  x  3.5)  +  (11  x  15) 

7.06  x  120~  =  3'49  inches< 

In  order  to  find  the  tension  we  must  change  our  original  formula  (/)  to  the 
following : 

(P  x  A  x  B)  -  (W2  x  B)  -  (W3  x  C)  =  T  x  L.  (j) 

From  this  we  obtain 

_  (P*AxB)-(W2xB)-(W3x  C)  ,M 

~L~ 

which  reads : 

RULE  95. — From  the  moment  of  the  steam  pressure  subtract  the  moment  of  the 
weight  of  valve  and  spindle,  also  subtract  the  moment  of  the  lever;  divide  the 
remainder  by  the  length  of  the  lever ;  the  quotient  will  be  the  tension  in  pounds. 

EXAMPLE  124. — The  steam  pressure  is  120  pounds  per  square  inch;  all  other 
dimensions  are  as  given  in  Example  121,  excepting  the  tension,  which  is  to  be  found. 

Substituting  the  numerical  values  for  the  symbols  in  formula  (k),  we  have 

(120  x  7.06  x  3.5)  -  (5  x  3.5)  -  (11  x  15) 

1  '-  og  -  =  72.27  pounds. 

For  finding  the  length  L  of  the  safety-valve  lever  when  all  other  dimensions  and 
weights  are  given,  we  have  the  following  formula : 

(P  x  A  x  B)  -  (W2  x  B)  -  (W3  x  C)  A 

L  =  -  ~~T~  ~' 

In  this  formula  we  have  taken  into  account  W3,  which  is  the  weight  of  the  lever, 
and  C,  which  is  the  distance  from  the  fulcrum  to  the  center  of  gravity  of  the  lever, 
and,  since  these  cannot  be  accurately  determined  unless  the  length  L  of  the  lever  is 
known,  we  are  compelled  to  assume  a  value  for  W3  x  C. 

Probably  under  these  circumstances  it  will  be  best  to  find  the  length  L  of  the 
lever  by  formula  (6),  Art.  373 ;  the  length  L  thus  found  will  be  somewhat  longer  than 
that  obtained  by  formula  (I). 

DOME  TOPS. 

375.  The  safety  valves  are  generally  attached  to  a  cast-iron  dome  top,  the  propor- 
tions of  which  are  shown  in  Figs.  554,  555 ;  the  relative  positions  of  the  safety  valves 
and  the  whistle  are  also  given. 

The  safety  valve,  such  as  is  shown  in  Fig.  553,  is  placed  in  the  opening  A,  and 
the  locked  safety  valve,  or  pop  valve,  is  placed  in  the  opening  B.  The  whistle  is 
screwed  into  the  central  hub  E,  and  stands  forward  of  the  safety  valves. 

When  the  ordinary  plain  safety  valve  is  used,  we  do  not  require  the  hubs  f,  f, 
shown  in  Fig.  555 ;  yet  these  are  generally  cast  to  the  dome  top  so  that  any  other 
safety  valve  can  readily  be  attached.  The  basin  C  is  simply  for  the  purpose  of 


MODERN   LOCOMOTIVE    CONSTRUCTION. 


381 


collecting  the  water  due  to  the 
condensation  of  steam  as  it 
flows  through  the  safety  valves 
and  whistle. 

Fig.  556  represents  a  plan 
of  the  dome  ring,  and  the  ar- 
rangements of  the  rivet  holes 
/•  /•,  and  the  stud  holes  s  s. 

Fig,  556A  represents,  on  a 
larger  scale,  the  joint  between 
the  dome  top  and  the  ring ;  this 
joint  is  a  ground  one.  F  repre- 
sents a  section  of  the  dome-top 
flange,  and  H  a  section  of  the 
ring.  Here  it  will  be  seen  that 
the  holes  s  for  the  studs  I  are 
not  drilled  through  the  ring; 
this  precaution  is  taken  to  pre- 
vent leakage  around  the  studs. 

Sometimes  the  rings  are 
placed  otitside  of  the  dome;  in 
such  cases  bolts  are  used  for 
fastening  the  dome  top  to  the 
ring. 

WHISTLE. 

376.  Fig.  557  shows  a  sec- 
tion of  the  whistle,  a  section  of  a 


Fig.  556 


Fig.SSS 


common  safety  valve,  and  a  sec- 
tion of  a  pop  safety  valve.  An- 
other section  of  the  lower  por- 
tion of  the  whistle  is  shown  in 
Fig.  559.  The  whistle  consists 
of  the  bell  0,  generally  made  of 
brass  ;  the  stem  P,  generally 
made  of  malleable  iron — some- 
times of  wrought-iron  ;  the 
brass  bowl  N  and  the  shank 
M  cast  in  one  piece,  the  brass 
disk  W,  the  brass  valve  R,  and 
the  wrought-iron  lever  T.  The 
valve  M  rests  against  the  coni- 
cal seat  formed  on  the  bottom 
of  the  shank.  The  valve  is 


MOltEUX  LOCOMOTIVE   CONSTRUVT1ON. 


MODERN  LOCOMOTIVE  CONSTRUCTION: 

made  with  guides  or  wings  e  e  extending  upwards  in  the  shank  for  a  considerable 
distance,  so  as  to  provide  for  a  pocket  for  the  screw  d,  which  prevents  the  valve 
from  turning,  and  from  dropping  into  the  boiler  when  the  lever  T  is  at  any  time 
taken  out;  and  also,  for  a  pocket  which  receives  the  end  of  the  lever  T;  the  height 
at  which  the  lever  T  must  be  placed  above  the  dome  is  generally  determined  by 
the  height  of  the  cab  above  which  the  lever  T  must  pass.  The  disk  W  is  held  in 
position  by  the  stem  P,  which  is  screwed  into  the  hub  cast  in  the  center  of  bowl. 
An  annular  opening  b  b  is  left  between  the  disk  W  and  the  inner  surface  of  the 
bowl.  The  upper  end  of  the  bell  0  is  tapped,  and  can  be  set  to  any  desired  height 
on  the  stem  P,  and  is  secured  in  its  position  by  the  jam-nut  U.  The  outer  diameter 
of  the  bell  must,  of  course,  always  be  a  little  larger  than  the  outer  diameter  of  the 
annular  opening  b  b.  The  lever  T  works  on  the  fulcrum  c,  and  it  must  move  the 
valve  E  downwards  when  a  communication  with  the  steam  space  in  the  boiler  is  to  be 
opened,  enabling  the  steam  to  flow  upwards  in  the  shank,  then  pass  through  openings 
a  a,  and  finally  flow  out  through  the  annular  opening  b  b,  striking  the  lower  end  of 
the  bell  0,  thereby  producing  either  a  deep  or  shrill  sound,  according  to  the  size 
and  proportions  of  the  whistle. 

Fig.  560  represents  part  of  the  plan  of  the  bowl,  and  a  section  through  the  valve 
and  fulcrum  c.  The  size  of  whistle  is  designated  by  the  outer  diameter  of  the  bell ; 
the  whistle  here  given  is  called  a  6-inch  whistle,  and  this  size  is  the  most  common  one, 
used  on  nearly  all  locomotives;  sometimes  a  4-inch  whistle  is  adopted  for  small 
locomotives  running  on  ordinary  surface  roads. 

Fig.  561  represents  another  whistle  lever,  B,  which  is  placed  in  the  cab.  The  rod 
marked  V  in  Figs.  561  and  559  represents  one  and  the  same  rod,  which  passes  through 
the  roof  of  the  cab  and  connects  the  lever  B  to  the  lever  T  (Fig. 
559).  The  lever  B  works  on  the  fulcrum  A,  whose  shank  a 
passes  through  the  roof  of  the  cab  and  is  fastened  there.  A 
plan  of  the  lever  B  is  shown  in  Fig.  5610.  This  arrangement 
of  whistle  is  often,  but  not  exclusively  used ;  indeed,  sometimes 
the  whistle  is  operated  simply  by  a  cord. 

CHIME  WHISTLE. 

377.  Fig.  562  represents  a  single  bell  chime  whistle ;  *  the 
peculiar  features  of  this  whistle  are  found  in  the  construction  of 
the  bell,  which  is  divided  into  three  compartments.  One  of 
these  compartments  extends  throughout  the  whole  length  of  the 
bell;  the  second  compartment  is  made  somewhat  shorter,  and 
the  third  still  shorter.  The  whistle  valve  need  not  differ  in  con- 
struction from  that  of  any  other  whistle,  and  can  be  made  to  suit 
the  taste  and  experience  of  the  designer.  This  whistle  produces  7'  "A  562 

three  distinct  tones,  which  harmonize  and  give  an  agreeable  musical  chord ;  and  when 
they  •  are  used  on  passenger  trains  exclusively,  serve  to  distinguish  the  latter  from 
freight  trains.  These  whistles  have  been  adopted  for  passenger  service  on  several 
railroads,  and  are  favorably  endorsed. 

*  Patented  and  made  by  the  Crosby  Steam  Gauge  and  Valve  Co.,  of  Boston. 


384  MODERN  LOCOMOTIVE   CONSTRUCTION. 


POP  SAFETY  VALVES. 

378.  In  Fig.  557  the  section  A  is  that  of  a  common  safety  valve,  which  has  been 
illustrated  in  Tig.  551.  B  is  a  section  of  a  pop  safety  valve,  first  introduced  on 
locomotives  by  George  W.  Eichardson.  The  openings  in  the  dome  top  for  both  safety 
valves  are  generally  3  inches  diameter,  but  for  the  pop  valve  the  opening  is  reduced 
by  the  brass  bushing  L,  making  the  opening  for  the  safety  valve  2J  inches  diameter, 
which  is  the  size  of  all  locomotive  pop  safety  valves.  The  brass  bushing  L  is  pressed 
tightly  into  the  dome  top.  On  the  top  of  this  bushing  a  conical  valve  seat  C  is 
formed,  on  which  the  valve  B  rests.  The  valve  is  made  with  wings  or  guides  extend- 
ing nearly  to  the  bottom  of  the  bushing.  In  the  center  of  the  valve  a  hole  of  consider- 
able depth  is  drilled,  and  on  the  bottom  of  this  hole  a  spindle  F  rests,  which  fits  very 
loosely  in  the  hole.  The  disk  G,  which  is  pressed  on  the  spindle,  supports  the  spring 
S;  the  upper  end  of  this  spring  acts  against  a  disk  cast  to  the  cross-bar  /,  usually 
made  of  brass.  On  account  of  showing  the  valve  in  section,  one-half  of  the  cross-bar 
is  cut  off,  and  the  remaining  half  is  shown  foreshortened.  A  full  view  of  the  cross- 
bar is  given  in  Fig.  557a.  The  upper  end  of  the  spindle  F  is  guided  by  the  cross-bar. 
Two  wrought-iron  studs,  H  H,  are  screwed  into  the  dome  top ;  one  of  these  only 
is  shown  in  Fig.  557.  The  nuts  on  these  studs  regulate  the  height  at  which  the 
cross-bar  is  to  be  placed  above  the  valve,  and  thereby  regulating  the  resistance  of 
the  spring.  The  set-screw  K  is  for  the  purpose  of  preventing  the  valve  from  lifting 
too  high. 

When  common  safety  valves  are  held  down  by  springs  arranged  as  shown  in  Fig. 
552,  a  difficulty  presents  itself,  namely,  that  the  resistance  of  the  spring  increases 
with  the  lift,  and  therefore  the  lift  will  be  insufficient,  and  will  not  be  as  great  as  that 
of  a  safety  valve  held  down  by  an  ordinary  weight ;  even  with  the  latter  arrangement, 
the  lift  of  the  valve  in  many  cases  is  less  than  desired.  The  pop  valve  is  designed  to 
overcome  this  difficulty,  and  therefore  the  diameter  of  the  upper  end  of  the  valve  B 
is  made  considerably  greater  than  the  diameter  of  the  opening,  and  a  groove  is  turned 
in  the  lower  face  of  this  enlarged  end  of  the  valve ;  another  groove  E  is  turned  in  the 
upper  face  of  the  bushing,  both  grooves  being  outside  of  the  valve  seat.  Now,  as  the 
valve  lifts  these  grooves  fill  with  steam,  causing  it  to  act  with  considerable  energy 
on  the  increased  surfaces  of  the  valve  and  seat,  enabling  it  to  overcome  to  some 
extent  the  increased  resistance  of  the  springs,  and  giving  it  a  lift  which  will  increase 
the  venting  capacity  of  the  pop  valve  to  more  than  double  that  of  the  ordinary  safety 
valve  loaded  with  a  weight. 

The  general  design  of  this  valve  and  its  seat  is  shown  much  plainer  in  Fig.  5571 ; 
but  the  valve  here  represented  has  an  additional  piece,  namely,  the  adjustable  screw 
ring  K.  The  purpose  of  this  ring  is  to  regulate  the  difference  between  the  blowing- 
off  pressure  and  that  at  the  time  of  the  closing  of  the  valve.  If  the  boiler  pressure  is 
reduced  too  much  before  the  valve  closes,  loosen  the  set-screws  L  L,  and  turn  the 
ring  up  a  notch  at  a  time ;  if  it  reduces  the  pressure  too  little,  turn  the  ring  down 
a  notch  at  a  time,  until  the  desired  pressure  is  reached,  and  then  turn  down  the  set- 
screws. 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


385 


Sometimes  solid  nickel  valve  seats  are  used,  and  the  spring  is  nickel-plated  so  as 
to  prevent  corrosion. 

The  valve  shown  in  Fig.  5576  represents  one  of  Richardson's  patent  open  locomo- 


Fig.  5576. 


Fig.  557c. 

tive  pops;  and  Fig.  557c  shows  an  encased  one.'  The 
casing  greatly  reduces  the  noise  of  the  escaping  steam, 
and  the  lock-up  device  prevents  tampering  with  the 
valve. 

Fig.  563  represents  a  Crosby  pop  safety  valve. 
In  this  design  the  valve  B  rests  upon  two  flat  annu- 
lar seats  V  V  and  W  W,  winch  lie  in  the  same  plane, 
and  form  a  part  of  the  shell  A  A.  This  shell  may  be 
said  to  consist  of  two  distinct  parts ;  namely,  the  in- 
ner cylindrical  chamber  C  C,  and  the  outer  cylinder 
A  A  ;  these  two  are  connected  by  the  hollow  arms  D  D 
radiating  horizontally,  and  allowing  the  steam  to  pass 
between  the  arms  and  act  against  that  portion  of  the 
lower  surface  of  the  valve  B  which  is  exposed  between 
the  annular  valve  seats.  The  valve  B  is  made  with 
wings  or  guides  X  X,  which  project  into  cylindrical 
chamber  C.  When  the  valve  is  raised  against  the  re- 
sistance of  the  spring  S,  the  steam  flows  over  the  inner 
annular  valve  seat  W,  passes  into  the  cylindrical 
chamber  C,  and  out  through  the  passages  E  E  formed 
in  the  arms  D  D.  The  flange  G  G  is  for  the  purpose  of 


386  MODERN  LOCOMOTIVE  CONSTRUCTION. 

turning  the  current  of  steam  upwards.  The  tension  of  the  spring  S  is  regulated  by  the 
screw  bolt  L.  The  casing  around  the  spring  sei'ves  for  a  frame  against  which  the  end 
of  the  spring  acts,  and  also  reduces  the  noise  of  the  escaping  steam.  In  this  valve  also, 
provisions  have  been  made  to  overcome  the  increasing  resistance  of  the  spring  when 
the  valve  is  raised  off  its  seat,  for  we  have  already  seen  that  when  the  valve  is  closed 
the  steam  acts  on  that  portion  of  the  lower  surface  of  the  valve  which  is  confined 
within  the  annular  valve  seats  V  and  TF;  and  now,  when  the  valve  is  raised  off  its 
seat,  the  whole  lower  surface  of  the  valve  is  exposed  to  the  action  of  the  steam. 

The  pressure  which  the  spring  8  in  this  valve  has  to  resist  when  the  valve  is 
closed  will  be  equal  to  the  area  of  the  lower  surface  of  the  valve,  which  is  confined 
within  the  outer  edge  of  the  annular  valve  seat  TF,  and  the  inner  edge  of  the  annular 
valve  seat  V,  multiplied  by  the  steam  pressure  per  square  inch. 

In  the  Richardson  safety  valve,  the  pressure  which  the  spring  has  to  resist  when  the 
valve  is  closed  will  be  equal  to  the  area  of  the  safety-valve  opening  multiplied  by  the 
steam  pressure  per  square  inch. 

HELICAL   SPEINGS. 

379.  Rules  relating  to  the  strength  of  helical  springs  —  such  as  are  used  for  safety 
valves  —  have  been  given  by  different  authorities  ;  some  of  these  rules  are  not  satis- 
factory, as  they  do  not  give  results  agreeing  with  practice.  The  best  we  have  seen  are 
given  in  "  Manual  of  Rules,  Tables  and  Data  for  Mechanical  Engineers,"  by  D.  K. 
Clark,  from  which  the  following  formulas  have  been  taken  : 


. 

in  which  : 

E  =  Compression  or  extension  of  one  coil,  in  inches. 

d  =  Diameter  from  center  to  center  of  steel  bar  composing  the  spring  in  inches. 

D  =  Diameter,  or  side  of  square  of  the  steel  bar,  of  which  the  spring  is  made,  in  six- 

teenths of  an  inch. 
C  =  A  constant,  which  from  experiments  made  may  be  taken  as  22  for  round  steel, 

and  30  for  square  steel. 
•w  =  Load  on  spring,  in  pounds. 

The  deflection  of  one  coil  is  to  be  multiplied  by  the  number  of  free  coils  to  obtain 
the  total  deflection  for  a  given  spring.  The  deflection  obtained  by  this  formula  for 
springs  made  of  f  square  steel  agrees  very  closely  with  practice.  But  for  springs 
made  of  steel  less  than  |  of  an  inch  square  we  would  suggest  to  increase  the  constant 
(30),  as  experiments  seem  to  indicate  that  steel  rolled  to  the  smaller  sizes  does  not 
deflect  as  much  proportionally.  For  steel  £  of  an  inch  square  a  constant  of  40  will 
give  better  results  than  a  constant  of  30. 

To  find  the  size  of  steel  for  a  given  diameter  of  spring  and  pressure,  we  have 
the  following  formulae,  also  taken  from  D.  K.  Clark's  work  : 

3  /  —      —  r- 

-  for  round  steel,  (2) 


D  =z  \/  -         -  for  square  steel.  (3) 

V      4.29 


MODERN  LOCOMOTIVE   CONSTRUCTION.  387 

The  letters  given  in  these  formulas  represent  the  same  quantities  as  given  for  the 
first  formula. 

EXAMPLE  125. — How  much  must  a  helical  spring  be  compressed  to  resist  a  steam 
pressure  of  120  pounds  per  square  inch  ?  Safety-valve  opening,  2 J  inches  diameter ; 
spring  made  of  g-iuch  square  steel ;  diameter  of  spring  from  center  to  center  of  coil, 
2  inches ;  6  free  coils. 

In  applying  formula  (1)  we  have  d3  =  2  x  2  x  2.  Total  pressure  on  the  valve  is 
equal  to  its  area  multiplied  by  the  steam  pressure  =  3.97  x  120  =  476.4  pounds  =  w. 
Steel  |  inch  square  =  -&,  hence  D4  =6x6x6x6.  And  for  the  constant  C  in  this 
example  we  shall  use  30.  Hence  we  have 

2  x  2  x  2  x  476.4      3811.2 
6x6x6  x  6  x  30  "  38880  * 

Here,  then,  the  compression  for  1  coil  is  .098  of  an  inch ;  consequently  for  6  free 
coils  the  compression  will  be  .098  x  6  =  .588,  or  nearly  ££  of  an  inch. 

EXAMPLE  126. — The  total  pressure  which  a  helical  spring  has  to  resist  is  476.4 
pounds ;  diameter  of  coil  from  center  to  center  of  steel  bar  composing  it  is  to  be  2 
inches ;  find  the  size  of  square  steel  required. 

According  to  formula  (3),  we  have 


3  / —  S  / — 

#  -  V  -  i=V"      7m  ^222  =  6.05  sixteenths  inch, 

4.—.'  I ._.' 

Sa7  A  =  §  of  an  inch  square  for  the  size  of  steel  bar. 


DOME  TOP  MADE  IN   TWO   PIECES. 

380.  In  Figs.  554,  555  we  have  shown  a  dome  top  made  in  one  piece,  but  fre- 
quently dome  tops  are  made  in  two  pieces.  In  the  latter  design,  one  piece  is  called 
the  dome  cover,  and  the  other  the  dome  ring. 

Fig.  564  represents  a  section,  and  Fig.  565  a  plan  of  the  dome  cover ;  Figs.  566 
and  5<J7  represent  the  section  and  plan  of  the  dome  ring.  The  dome  ring  is  riveted  to 
the  dome  sheet ;  the  manner  of  fastening  the  dome  cover  to  the  ring  and  also  their 
dimensions  are  plainly  indicated  in  the  illustrations. 


PUMPS. 

381.  In  nearly  all  modern  locomotives  the  water  is  supplied  to  the  boiler  by 
injectors;  in  a  few  engines  the  pump,  which  in  former  years  was  used  exclusively, 
still  holds  its  place,  and  even  in  these  few  cases  we  sometimes  find  only  one  pump 
which  is  placed  on  one  side  of  the  engine,  and  an  injector  on  the  opposite  side.  Pumps 
are  often  worked  directly  from  the  crosshead;  these  ai-e  called  full-stroke  pumps. 
Sometimes  they  are  worked  from  an  eccentric  fastened  to  one  of  the  driving  axles, 
and  others  are  worked  from  a  small  crank  attached  to  the  crank-pin.  The  stroke  of 
the  latter  class  of  pumps  is  generally  less  than  that  of  the  piston,  and  therefore  they 
are  called  short-stroke  pumps. 


388 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


Figs.  568,  569,  and  570  represent  a  full-stroke  pump  designed  for  an  eight-wheeled 
passenger  engine.  Fig.  568  represents  a  section;  Fig.  569  a  plan;  Fig.  570  an  end 
view  of  the  pump,  and  Fig.  570a  a  section  through  the  lower  stuffing-box  and  gland. 
The  pump  consists  of  the  barrel  B,  the  lower  air-chamber  A,  the  upper  air-chamber 
C,  the  valves  L  L,  the  valve  seats  1 1,  and  the  cages  K  K.  The  valves  are  cylindrical 


Fig.  565 


Fig.  567 

in  form,  and  rest  on  the  seats  I J,  the  latter  resembling  ordinary  disks.  The  upper 
valve  seat  rests  on  the  face  of  the  flange  of  the  barrel,  and  the  lower  valve  seat  rests 
on  the  face  of  the  flange  on  the  air-chamber.  The  cages  rest  on  the  valve  seats,  the 
whole  being  clamped  together  and  held  in  position  by  the  bolts  through  the  flanges  of 
the  air-chambers,  and  those  of  the  barrel.  The  joints  between  the  flanges,  valve  seats, 
and  cages  are  ground  joints.  An  enlarged  section  and  plan  of  the  cage  and  valve  seat 
are  shown  in  Figs.  571  and  572.  The  purpose  of  the  cage  is  to  guide  the  valve,  and 
prevent  it  from  lifting  too  high. 

The  barrel  and  air-chambers  are  often  made  of  brass ;  sometimes  of  cast-iron ;  the 
valves,  seats,  and  cages  are  always  made  of  brass. 

The  stuffing-box  on  the  lower  air-chamber  receives  the  end  of  the  suction  pipe,  and 
a  water-tight  connection  is  made  by  means  of  the  hemp  packing  in  the  stuffing-box  E. 
The  stuffing-box  on  the  upper  chamber  receives  the  end  of  the  delivery  pipe,  and  here, 
also,  a  water-tight  connection  is  obtained  by  means  of  the  hemp  packing  in  the  stuffing- 
box  G.  The  suction  pipe  extends  to  the  rear  end  of  the  engine,  where  it  is  connected 
by  a  rubber  hose  to  the  tank.  Since  the  suction  pipe  is  comparatively  a  straight  pipe,  it 
is  often  made  of  iron ;  sometimes  of  brass.  The  delivery  pipe  extends  to  the  check  valve 
on  the  boiler,  and  because  the  delivery  pipes  have  to  be  bent,  they  are  made  of  copper, 


MODERX  LOCOMOTIVE   CONSTRUCTION. 


389 


The  manner  of  fastening  the  pump  to  the  frame  F  is  plainly  shown  in  Fig.  569. 

Fig.  573  represents  the  plunger ;  it  is  made  of  a  wrought-iron  bar,  with  a  tapered 
shank  «  turned  at  one  end.  This  shank  fits  in  a  lug  either  cast  or  bolted  to  the  crosshead. 

In  a  pump  of  the  size  here  shown,  the  lift  of  the  lower  valve — sometimes  called  the 
suction  valve — is  one-quarter  of  an  inch ;  and  the  lift  of  the  upper  valve — sometimes 


390  MODERN   LOCOMOTIVE   CONSTRUCTION. 

called  the  delivery  valve — is  •£§  of  an  inch.  The  successful  working  of  the  pumps 
depends  to  a  considerable  extent  on  the  correct  amount  of  lift,  hence  the  necessity  of 
exercising  great  care  in  determining  the  lift ;  too  much  lift  will  prevent  a  quick  closing 
of  the  valve ;  it  will  also  cause  it  to  pound,  thereby  ruining  the  valve  seat. 

382.  The  air-chambers  are  for  the  purpose  of  relieving  the  pump  and  its  pipes  from 
sudden  shocks,  which  are  liable  to  occur  by  the  rapid  motion  of  the  pump  plunger. 
For  instance,  as  soon  as  the  water  rises  above  the  top  of  the  opening  o  in  the  upper  air- 
chamber,  the  air  in  the  latter  will  be  compi-essed,  and  act  as  a  cushion,  thereby  reduc- 
ing the  intensity  of  a  shock.  The  lower  air-chamber  is  arranged  somewhat  differently ;  a 
pipe  M  extends  from  the  top  to  within  a  short  distance  from  the  bottom  of  the  chamber, 
so  as  to  leave  an  annular  air  space.     This  pipe  M  is  called  a  dip  pipe.    As  soon  as  the 
water  reaches  the  lower  end  p  of  the  dip  pipe,  the  air  in  the  annular  space  is  com- 
pressed, and  has  a  similar  effect  as  the  air  in  the  upper  chamber.     It  must  readily  be 
perceived  that  when  the  lower  air-chamber  does  not  contain  a  dip  pipe,  the  chamber 
will  soon  be  rendered  useless,  because  the  air  in  it  will  be  replaced  by  water. 

In  many  locomotive  pumps  the  water  is  discharged  through  the  top  of  the  upper 
air-chamber,  instead  of  through  its  side,  as  shown  in  our  illustrations.  In  cases  of  this 
kind  the  upper  chamber  must  be  arranged  similar  to  the  lower  one ;  that  is  to  say,  it 
must  be  provided  with  a  dip  pipe  extending  from  the  top  of  the  chamber  to  within  a 
short  distance  from  its  lower  end ;  otherwise  the  chamber  will  soon  be  filled  with  water, 
and  rendered  useless. 

The  air  capacity  above  the  water  line  in  the  upper  chamber  should  be  equal 
to  the  total  displacement  of  the  pump  plunger;  in  fact,  one  and  a  half  times  the 
displacement  will  give  better  results.  Thus,  for  instance:  Let  the  plunger  be  2 
inches  in  diameter,  and  24  inches  stroke.  The  total  displacement  of  the  plunger 
is  found  by  multiplying  its  cross-sectional  area  by  the  stroke.  Now,  the  area  of  a 
cross-section  2  inches  in  diameter  is  3.14  square  inches,  hence  the  total  displace- 
ment will  be  3.14  x  24  =  75.36  cubic  inches.  Consequently  the  air  capacity  of  the 
upper  chamber  should  not  be  less  than  75.36  cubic  inches,  and  75.36  cubic  inches 

-I-  — ^—  =  113.04  cubic  inches  will  be  better. 

The  inner  diameter  of  the  dip  pipe  must  be  equal  to  the  inner  diameter  of  the  suc- 
tion or  feed  pipe ;  and  since  the  air  capacity  in  the  annular  space  should  be  equal  to 
the  air  capacity  of  a  plain  chamber,  it  follows  that  many  air-chambers  provided  with 
dip  pipes  have  larger  outer  dimensions  than  plain  chambers. 

The  air  capacity  of  the  lower  air-chamber  can  always  be  made  a  little  less  than 
that  of  the  upper  one. 

383.  The  small  pet-cock  D,  shown  in  Fig.  570,  is  for  the  purpose  of  ascertaining 
whether  the  pump  is  feeding  water  into  the  boiler. 

Fig.  574  represents  different  views  and  details  of  the  pet-cock  (marked  D  in  Fig.  570) 
and  its  fittings.  The  lower  end  of  the  rod  A  fits  in  a  square  pocket  cast  into  the  pet- 
cock  plug.  In  many  engines  this  rod  extends  a  little  above  the  running  board ;  the 
lever  J5,  attached  to  the  top  of  the  rod  A,  is  connected  to  the  reach-rod  C,  which  leads 
into  the  cab,  enabling  the  engineer  to  open  or  close  the  cock,  and  thus  determine 
whether  the  pump  is  feeding  water  into  the  boiler. 


MODERX  LOCOMOTirK   CONSTRUCTION. 


391 


CHECK   VALVES. 


384.  Figs.  575,  570,  577  represent  the  check  valve.  Its  purpose  is  to  prevent  the 
water  in  the  boiler  from  flowing  back  into  the  pump,  and  also  to  prevent  the  water  from 
flowing  out  of  the  boiler  and  doing  serious  damage  in  case  of  accidents  to  the  pump  or 
del  ivory  pipe.  The  check  valve  consists  of  a  brass  or  cast-iron  case  A,  the  valve  L,  the 


Pet  Cock  and  Cutmrrlions 

The  name  for  all  Cyh./row  I'i  upwanll 


Check  Valve 

The  snmefor  all  Cl/li./rotH 


Vole 

Sirrf  Ileadt  outtidc 

Jiieelal  uaUe  <tf  Shell 


cage  K  and  the  lower  section  or  bottom  B ;  the 
upper  face  of  this  section  B  forms  the  valve 
seat,  and  the  lower  part  of  this  section  is  con- 
nected to  the  delivery  pipe  by  the  coupling 
nut  C.  The  flange  D  is  riveted  to  the  boiler 
shell;  the  hub  of  this  flange  extends  to  the 
outside  of  the  lagging  E,  and  forms  a  ball  joint  with  the  check  valve.  The  manner  of 
fastening  the  check  valve  to  the  boiler  is  plainly  shown.  F  is  simply  a  false  cover ;  it 
is  not  shown  in  Fig.  577. 

The  check  valve  should  be  placed  towards  the  front  of  the  boiler ;  it  is  usually 


392 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


placed  at  a  distance  of  Ij  to  2  feet  from,  and  in  the  rear  of,  the  front  flue  sheet,  so  as 
to  feed  the  water  into  the  coolest  part  of  the  boiler,  and  also  prevent  as  much  as  pos- 
sible the  collections  of  any  impurities,  which  may  enter  into  the  boiler  with  the  water, 
around  the  hot  sheets  in  the  furnace.  These  impurities  generally  collect  under  the 
check- valve  opening  in  the  boiler  in  the  form  of  mud,  and  pile  up  towards  the  front  flue 
sheet.  If  the  check  valve  is  connected  to  the  back  end  of  the  boiler,  as  has  been  done 
in  a  few  instances,  the  impurities  will  collect  in  the  water  space  around  the  fire-box, 
causing  the  sheets  to  be  burnt. 

It  is  also  important  that  the  valve  L  should  stand  in  a  vertical  position,  conse- 
quently the  center  of  the  check-valve  opening  in  the  boiler  is  generally  placed  in  a 
horizontal  plane  passing  through  the  axis  of  the  barrel  of  the  boiler. 

For  injectors,  the  check  valves  should  be  placed  in  the  same  position  as  those  for 
pumps. 

In  many  engines  we  find  the  check  valve  attached  to  the  boiler  in  a  manner  differ- 
ent from  that  shown  in  our  illustrations ;  instead  of  the  brass  flange  7),  a  wrought-irou 
plate  (reinforcing  plate),  about  £  inch  thick,  is  riveted  to  the  inside  of  the  boiler,  and 
both  the  boiler  sheet  and  reinforcing  plate  are  drilled  and  tapped  to  receive  the  shank 
of  the  check  valve.  Nearly  all  the  check  valves  for  injectors  are  attached  to  the  boiler 
in  this  manner. 

FEED-COCKS. 

385.  Figs.  578  to  581  inclusive  represent  different  views  of  the  feed-cock;  its 
construction  is  plainly  shown,  and  does  not  need  a  description.  The  kind  of  feed- 


Fved.  Cock     Tlte  tamp  for  all.  Cyl*.  from  12'tijucards. 


cock  here  shown  is  preferable  to  one  whose  plug  extends  through  the  bottom  of  the 
casing,  and  adjusted  by  a  nut  at  the  bottom;  feed-cocks  of  the  latter  class  are  liable 
to  leak ;  or,  when  adjusted  to  prevent  leaking,  will  often  require  too  much  power  to 
open  or  close  them. 


MODERN  LOCOMOTIYE   COXSTKVCTIOlf. 


393 


JTi<j.  .W.7<| 


feed  Cock  Qtuidrnnt  and  Handle 

The  tame/or  all  Cyla.  front  I21  upuj't 


The  inside  of  the  suc- 
tion hose,  running  from 
engine  to  tender,  is  often 
lined  with  brass  or  iron 
wire,  wound  spirally  to 
prevent  it  from  buckling. 

A  brass  sleeve  is  in- 
serted in  each  end,  thereby 
providing  means  for  coup- 
ling the  hose  to  engine  and 
tender. 

When  the  tender  is  to 
be  uncoupled  from  the  en- 
gine, the  hose  is  generally 
detached  from  the  feed- 
cock;  to  do  so  conven- 
iently, without  using  a 
wrench,  a  wing  nut  C  is 
used  —  of  which  another 
view  is  shown  in  Fig.  578« 
— at  this  end  of  the  hose ; 
the  other  end  of  the  hose 
is  sometimes  attached  by 
means  of  a  spanner  nut 
to  the  goose-neck,  as  shown 
in  Fig.  582,  but  the  best 


practice  is  to  use  a  wing  nut  at 

both  ends   of  the  hose,  so   that        Ft/a' ' 

either  of  its  ends  can  be  coupled  to  the  feed-cock. 


GOOSE-NECK. 


386.  The  goose-neck  is  shown  in  Figs.  582,  583,  584 ;  the 
spanner  nut  just  referred  to  is  marki'd  /.'  in  Fig.  5S2,  and  the 
sleeve  which  is  to  be  inserted  and  fastened  to  the  end  of  the 


394  MODERN  LOCOMOTIVE   CONSTRUCTION: 

hose  is  also  shown  in  this  figure.  The  goose-neck  is  made  of  cast-iron,  and  is  bolted 
to  the  bottom  of  the  tank.  The  conical  counterbore  a  forms  the  valve  seat.  The 
manner  of  operating  this  valve  will  be  shown  later. 

FEED-PIPE  HANGERS  AND  CONNECTIONS. — POSITION   OF  PUMPS. 

387.  The  feed-cock  is  held  in  position  by  clamping  its  branch  B  in  the  feed-pipe 
hanger  D,  which  is  shown  in  Fig.  585.     The  shank  of  this  hanger  takes  the  place  of  a 
bolt  for  bolting  the  frame,  foot  plate,  and  cab  bracket  together. 

The  socket  in  the  feed-cock  rod  E  fits  the  square  end  of  the  feed-cock  plug  A ; 
the  other  end  of  the  rod  E  works  in  a  bearing  drilled  into  the  frame.  The  crank  or 
lever  F  on  the  rod  E  is  connected  to  the  crank  G  on  the  quadrant  rod  K ;  the  upper 
end  of  this  rod  works  in  a  bearing  drilled  through  the  quadrant  7,  of  which  other 
views  are  shown  in  Figs.  585«  and  5856. 

The  quadrant  /  is  bolted  to  the  back  end  of  the  boiler  in  such  a  position  as  to 
bring  the  handle  H  within  easy  reach  of  the  engineer,  enabling  him  to  open,  close,  and 
regulate  the  amount  of  opening  in  the  feed-cock.  The  pointer  J  fastened  to  the  upper 
end  of  the  quadrant  rod  indicates  on  the  quadrant  the  amount  of  opening  in  the  feed- 
cock. 

In  order  to  obtain  a  close  regulation  of  feed- water,  the  opening  in  the  feed-cock 
plug  may  be  made  square,  with  a  diagonal  of  the  square  coinciding  with  the  axis  of 
the  plug ;  with  this  arrangement  a  closer  regulation  can  be  obtained  in  that  position 
of  the  feed-cock  plug  where  close  regulation  is  most  desirable. 

For  the  majority  of  injectors  at  present  in  use,  the  feed-cock  is  not  required,  and 
consequently  the  quadi-ant  and  rod  are  also  abolished;  but  non-lifting  injectors  do 
require  a  feed-cock ;  for  these,  a  feed-cock  with  a  square  opening,  as  described  above, 
is  recommended ;  the  opening  is  generally  £  of  an  inch  square. 

The  feed-pipe  hanger  D  is,  of  course,  retained  in  all  engines,  and  for  injectors 
this  hanger  takes  hold  of  a  brass  sleeve ;  the  rear  end  of  this  sleeve  is  threaded  like 
the  rear  branch  of  the  feed-cock,  and  the  front  end  of  the  sleeve  is  formed  like  the 
front  branch  of  the  feed-cock,  and  brazed  to  the  feed-pipe,  if  the  latter  is  made  of 
copper ;  iron  pipes  are  screwed  into  the  sleeve.  The  feed-pipe  hanger  shown  in  Fig. 
585  is  an  expensive  one  to  make,  hence  in  many  engines  we  find  the  hangers  made  as 
shown  in  Fig.  586.  The  rod  A  is  often  made  of  square  wrought-iron ;  the  shank  b, 
turned  at  the  end,  holds  the  cast-iron  clamps  c,  c,  which  support  the  sleeve  above 
referred  to.  The  cast-iron  clamps  are  usually  If  inches  wide. 

For  cold  climates  a  heater  pipe  about  f  inch  inside  diameter  is  attached  to  the 
suction  pipe,  enabling  the  engineer  to  blow  steam  into  the  latter,  to  prevent  freezing. 

388.  The  pumps  can  be  attached  directly  to  the  frames,  as  illustrated  in  Fig.  569, 
only  in  eight-wheeled  passenger  engines.     In  Mogul,  ten- wheeled,  and  consolidation 
engines  there  is  no  room  to  place  them  in  similar  positions,  therefore,  in  the  latter 
classes  of  engines,  we  often  find  them  attached  to  the  guides.    Figs.  587  and  588  show 
a  pump  designed  for  fastening  it  to  the  guide  and  guide-yoke;  the  only  difference 
between  this  pump  and  the  one  previously  shown  is  in  the  position  of  the  air-chambers, 
and  the  position  and  design  of  the  lugs.    The  type  of  guide  for  which  this  pump  was 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


395 


designed  is  shown  in  Fig.  241.     The  lug  near  the  stuffing-box  (Fig.  588)  is  bolted  to 
the  top  of  the  guide,  and  the  lug  near  the  air-chamber  is  bolted  to  the  guide-yoke. 

The  illustrations  (Figs.  587,  588)  show  the  air-chambers  placed  in  the  neighbor- 
hood of  the  center  of  the  barrel ;  the  object  aimed  at  in  this  design  is  to  concentrate 
the  weight  of  the  pump  as  much  as  possible  near  one  of  the  points  of  support.  As  far 
as  the  distribution  of  weight  is 
concerned,  the  design  is  correct ; 
but  in  placing  the  air-chambers 
at  or  near  the  center  of  the  bar- 
rel, the  direction  of  the  flow  of 
water  is  interfered  with,  which 
is  an  objection.  The  proper 
place  for  the  air-chambers,  in 
the  class  of  pumps  here  shown, 
is  at  the  end  of  the  barrel ;  and 
although  they  cannot  always  be 
placed  there,  it  should  be  re- 
membered that  any  change  in 
the  direction  of  the  flow  of  water, 
such,  for  instance,  as  is  caused 
by  placing  the  air-chamber  in 
the  center  of  the  barrel,  has  a 

tendency  to  reduce  the  efficiency  of  'the  pump,  and  this  should  be  avoided  as  much 
as  possible. 

It  may  also  be  well  to  remark  here  that  in  all  pumps  the  suction  valve  should  be 
placed  as  near  as  possible  to  the  plunger,  leaving  only  a  sufficient  water-way  between 
them.  Necessary  changes  in  the  form  and  size  of  the  water  passages  should  be  made 
gradually ;  sudden  enlargements  and  contraction  should  always  be  avoided. 


Fig.  SSS 


SHORT-STROKE   PUMPS. 

389.  We  have  previously  referred  to  short-stroke  pumps ;  sections  and  other  views 
of  this  class  of  pumps  are  given  in  Figs.  589  to  592.  These  pumps  are  placed  under 
the  boiler,  and  arc  bolted  to  a  cross-brace  A  extending  from  frame  to  frame.  The 
plunger  D  is  worked  from  an  eccentric  placed  on  one  of  the  driving  axles.  Fig.  593 
shows  separate  views  of  the  pump-rod  jaw  C',  to  which  the  pump-rod  D  is  connected. 
The  valves  and  cages  in  this  pump  arc  of  the  same  design  as  those  in  the  full-stroke 
pumps  previously  illustrated.  This  pump  is  designed  for  a  12-inch  cylinder. 

Figs.  594  to  596  represent  another  short-stroke  pump.  The  pump  is  fastened 
to  the  draw-bar,  and  is  worked  fi-om  a  pin  attached  to  one  of  the  crank-pins ;  hence 
the  pump  is  placed  outside  of  the  driving  wheels.  It  will  be  noticed  that  the  valves  L 
and  cages  K  are  made  somewhat  different  from  those  previously  illustrated;  this 
difference  is  not  due  to  any  particular  necessity,  but  is  simply  a  matter  of  choice. 


396 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


SIZE   OF   PUMPS. 

390.  To  find  the  size  of  pump  required,  we  should  estimate  the  consumption  of 
steam,  and  this  may  be  approximately  obtained  by  calculating  the  weight  of  steam  in 
the  cylinder  from  its  pressure  and  volume  at  the  time  of  its  release.  In  locomotives, 
the  average  pressure  of  the  steam  at  the  time  of  its  release  does  not  vary  much,  and 
therefore  we  may  make  the  pump  capacity  directly  proportional  to  the  cylinder 


(^M~£j-\ 

~= 

r  l%~'i 

I  -s*^j  
h  !s«';» 

1  \ 

n 
'%* 

^,,4. 


fia.  S91.. 


59O. 


capacity,  without  finding  the  amount  of  steam  consumed.  Since  the  stroke  of  a  full- 
stroke  pump  is  equal  to  that  of  the  piston  in  the  steam  cylinder,  we  may  further 
simplify  the  computation  and  make  the  cross-sectional  area  of  the  plunger  directly 
proportional  to  the  cross-sectional  area  of  the  cylinder.  Practice  indicates  that  for 
locomotives  having  two  pumps,  good  results  will  be  obtained  by  making  the  cross- 
sectional  area  of  each  plunger  equal  to  7\,  of  the  cross-sectional  area  of  one  cylinder. 
Tliis  proportion  is  not  strictly  adhered  to  by  the  different  builders ;  indeed,  many 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


397 


builders  use  the  same  size  of  pump  for  two  or  three  different  sizes  of  cylinders ;  for 
instance,  the  pump  illustrated  in  Fig.  568  is  used  for  cylinders  16,  17,  and  18  inches  in 
diameter.  But  the  proportion  here  given  we  believe  to  be  a  good  one  to  adopt  in 
designing  a  full-stroke  locomotive  pump.  Hence  the  following : 

RULE  96. — Divide  the  cross-sectional  area  of  the  cylinder  by  72 ;  the  quotient  will 


_ fl^ 


be  the  cross-sectional  area  of  the  pump  plunger  for  locomotives  in  which  two  full- 
stroke  pumps  are  to  be  used. 

EXAMPLE  127. — What  should  !><•  the  diameter  of  the   pump  plunger  for  a  full- 
stroke  pump  in  a  locomotive  having  cylinders  17  inches  diameter ! 


398  MODERN  LOCOMOTIVE   CONSTRUCTION. 

The  cross-sectional  area  of  a  cylinder  17  inches  diameter  is  226.98  square  inches, 

,  226.98 
and      _„        =  3.15  square  inches  for  the  cross-sectional  area  of  the  plunger;    the 

corresponding  diameter  is  2  inches  (very  nearly),  hence  the  diameter  of  the  plunger 
should  be  2  inches. 

Let  us  take  another  example.  Find  the  diameter  of  the  pump  plungers  for  a  full- 
stroke  pump  in  a  locomotive  having  cylinders  10  inches  diameter.  Here  we  have 

78.54 
„,,,      =  1.09  square  inches  for  the  cross-sectional  area  of  the  pump  plunger,  and  the 

corresponding  diameter  is  1^,  nearly. 

In  a  similar  manner  we  find  that  the  diameter  of  the  pump  plungers  for  full- 
stroke  pumps  for  cylinders  20  inches  in  diameter  will  he  nearly  2f  inches.  These 
results  agree  very  closely  with  the  average  locomotive  practice. 

The  capacity  of  a  full-stroke  pump  designed  by  the  foregoing  rule  will  be  ^  of 
the  capacity  of  one  steam  cylinder;  and  short-stroke  pumps  should  have  the  same 
capacity.  We  may  therefore  establish  the  following : 

BULE  97. — Multiply  the  cross-sectional  area  of  the  cylinder  in  square  inches  by 
the  length  of  stroke  of  piston  in  inches,  and  divide  the  product  by  72 ;  the  quotient 
will  be  the  capacity  of  the  pump  in  cubic  inches.  Dividing  this  capacity  by  the  length 
of  stroke  of  plunger  in  inches,  we  obtain  the  cross-sectional  area  of  the  plunger  in 
square  inches ;  or,  dividing  the  capacity  of  the  pump  by  the  cross-sectional  area  of  the 
plunger,  we  obtain  the  stroke  of  plunger. 

EXAMPLE  128. — Find  the  diameter  of  the  pump  plunger ;  its  stroke  is  7  inches ; 
the  diameter  of  the  steam  cylinder  is  18  inches ;  stroke,  24  inches. 

The  cross-sectional  area  of  the  steam  cylinder  is  254.47  square  inches ;  hence  its 
capacity  is 

254.47  x  24  =  6107.28  cubic  inches, 
and 

6107.28 
— ^ —    =  84.82  cubic  inches 

for  the  capacity  of  the  pump. 

Dividing  the  capacity  of  the  pump  by  the  stroke  of  plunger,  we  have 

84.82 

—=—  =  12.11  square  inches 

for  the  cross-sectional  area  of  the  plunger ;  and  its  diameter  will  be  3ff  inches  nearly ; 
hence  we  may  say  that  the  diameter  of  the  plunger  should  be  4  inches. 

Suppose  that  in  the  foregoing  example  the  diameter — i  inches — of  the  plunger 
had  been  given  instead  of  the  stroke,  and  that  it  is  required  to  find  the  stroke  of  the 
plunger.  Under  these  conditions  we  proceed  in  the  following  manner : 

We  first  find  the  cross-sectional  area  of  a  plunger  4  inches  diameter,  which  is 

84  8^ 
12.56  square  inches;  now,  dividing  the  pump  capacity  by  12.56,  we  have  ;.., '    .   =  6.75 

inches,  which  is  the  stroke  of  the  pump  plunger.  These  results  agree  closely  with  the 
size  of  the  pump  illustrated  in  Fig.  594,  which  is  an  exact  copy  of  a  working  drawing ; 


MODERN  LOCOMOTJfE  CONSTRUCTION.  399 

the  diameter  of  the  plunger  is  4  inches ;  stroke,  7  inches ;  this  pump  was  designed  for 
18  x  24  inch  cylinder. 

EXAMPLE  129. — Find  the  diameter  of  a  short-stroke  pump  plunger  whose  stroke 
is  3£  inches,  the  locomotive  having  cylinders  12  inches  in  diameter,  stroke,  20  inches. 

The  capacity  of  the  steam  cylinder  is  113.10  x  20  =  2262  cubic  inches.     That  of 

the  pump  should  be 

2262 

—=n~  =  31.41  cubic  inches ; 

and  the  cross-sectional  area  of  the  plunger  will  be 

31.41 

0  g    =  8.97  square  inches ; 
o.O 

and  its  diameter  will  be  3|  inches,  nearly.  Here,  again,  we  have  results  agreeing 
closely  with  dimension  in  Fig.  589,  which  is  also  a  copy  of  a  working  drawing. 

Of  course  the  fact  of  using  only  one  pump  and  an  injector  for  the  same  locomotive 
will  not  make  any  difference  in  the  sizes  of  pumps  found  by  the  foregoing  rules. 
When  no  injector  is  used,  two  pumps  of  the  sizes  here  given  will  be  required  for  each 
engine. 


CHAPTER    IX. 

SPRING   GEAR. 

391.  Fig  599  represents  the  spring  gear  for  the  driving  wheels  under  an  eight- 
wheeled  passenger  locomotive,  with  cylinders  17  inches  diameter  and  24  inches 
stroke.  Of  course  in  this  class  of  engines  we  have  only  two  driving  axles ;  these 
are  marked  F,  F2.  The  spring  saddles,  their  position  on  the  driving-boxes,  and  the 
manner  of  placing  the  springs  on  the  saddles,  have  been  illustrated  in  Figs.  303,  304. 
In  the  illustrations  before  us  we  have  only  shown  the  center  lines  8,  S2  of  the  spring 
saddles.  This  design  of  spring  gear  is  probably  the  most  common  one  used  for  eight- 
wheeled  engines  4  feet  8£  inches  gauge ;  a  few  builders  make  changes  in  the  minor 
details,  but  the  design  as  a  whole  remains  the  same.  A,  B,  and  C  are  the  spring 
hangers ;  sepai-ate  views  of  the  hanger  A  are  shown  in  Fig.  G03,  of  B,  in  Fig.  604, 
and  of  C,  in  Fig.  605.  Figs.  601,  602  show  separate  views  of  the  equalizing  lever  E, 
and  Fig.  600  shows  separate  views  of  the  lever  fulcrum  D ;  all  of  these  are  made  of 
wrought-iron. 

In  this  design  the  lever  fulcrum  is  bolted  to  the  frame  with  two  bolts,  but  we 
believe  it  to  be  better  practice  to  fasten  it  with  four  bolts  whenever  the  design 
of  engine  will  admit  that  number.  The  central  slot  in  the  equalizing  lever  is  cut 
sufficiently  long  to  allow  the  lever  to  vibrate  around  the  gib  or  key  through  the 
lever  fulcrum.  The  equalizing  lever  is  connected  to  the  springs  by  the  link  hangers 
B,  B2  passed  through  the  slots  in  the  springs,  and  in  the  ends  of  the  equalizing  lever ; 
the  outer  ends  of  the  springs  are  connected  to  the  frame  by  the  link  hangers  A  and  C; 
the  springs  and  equalizing  lever  form  a  system  of  levers  free  to  turn  about  their 
respective  fulcrums. 

Fig.  606  represents  the  driving-wheel  spring  gear  for  a  ten-wheeled  engine.  The 
only  difference  between  this  spring  gear  and  the  one  shown  in  Fig.  599  is  that  the 
former  has  an  additional  spring  and  equalizing  lever  on  each  side  of  the  engine. 

These  springs  support  the  greatest  portion  of  the  weight  of  the  engine;  a 
smaller  portion  is  supported  by  the  trucks ;  the  axles,  wheels,  driving-boxes,  spring 
saddles,  and  springs,  the  side-rods,  a  part  of  the  main-rods,  part  of  the  eccentric- 
rods  and  the  eccentrics  are  directly  supported  by  the  track ;  and  these  weights  sub- 
tracted from  the  weight  on  the  driving  wheels  will  give  the  load  on  the  driving-wheel 
springs. 

The  purpose  of  the  equalizing  levers  is  to  distribute  the  weight  equally  on  the 
driving  axles,  also  to  reduce  the  effects  of  shocks  caused  by  the  rails,  and  to  allow  the 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


401 


wheels  to  adjust  themselves  readily  to  any  unevenness  in  the  track  without  throwing 
an  undue  strain  on  the  frames  and  other  parts  of  the  locomotive.    It  will  readily  be 


seen  that  when  one  wheel  receives,  through  the  unevenness  of  the  road,  a  shock,  it  is 
immediately  transferred  to  the  spring;  a  part  of  this  shock  will  be  transferred  through 
the  outer  spring  hanger  to  the  frame,  and  the  other  part  will  be  transferred  to  the 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


equalizing  lever,  and  from  thence  thrown  on  the  next 
spring;  or,  again,  if  one  spring  becomes  momentarily 
burdened  by  an  oscillation  of  the  engine,  the  equalizing 
lever  will  immediately  transfer  a  part  of  this  load  to  the 
next  wheel,  thereby  distributing  the  effects  of  the  shock 
among  all  the  axles.  If,  on  the  other  hand,  an  equalizing 
lever  is  not  used,  the  whole  effect  of  the  shock  must  be 
resisted  by  one  spring,  and  this  spring  must  be  made 
strong  enough  to  enable  it  to  do  so ;  consequently  the 
spring  must  be  made  heavier  than  it  need  to  be  with  an 
equalizing  lever ;  but  making  the  spring  heavier  will  also 
reduce  its  elasticity,  which,  in  turn,  will  greatly  increase 
the  danger  of  breaking  the  axle,  saddle,  spring  hangers, 
and  spring ;  heavy  springs  will  also  make  the  riding  on 
the  engine  very  uncomfortable,  and,  in  some  cases,  to 
such  a  degree  as  to  be  almost  unbearable.  From  the 
foregoing  remarks  we  can  readily  conceive  the  advan- 
tages gained  by  the  use  of  equalizing  levers. 

In  eight-wheeled  passenger  engines,  and  also  in  ten- 
wheeled  engines,  all  the  driving-wheel  springs  are  con- 
nected by  equalizing  levers.  But  we  will  presently  see 
that  for  Mogul  and  consolidation  engines  it  is  not  practic- 
able to  connect  all  the  driving-wheel  springs  in  this  way. 

In  determining  the  thickness  and  depth  of  the  equal- 
izing lever  we  generally  find  that  the  design  of  the  engine 
will  establish  the  thickness.  For  instance,  in  many  engines 
the  pads  which  connect  the  boiler  to  the  frame  are  placed 
at  each  side  of  the  lever  fulcrum  D,  and  the  nuts  which 
fasten  the  pads  to  the  boiler  will  be  along  one  side  of  the 
equalizing  lever.  There  should  be  at  least  £  of  an  inch 
clearance  between  these;  and  since  the  levers  are  gen- 
erally made  of  a  symmetrical  form,  it  follows  that  the 
design  and  position  of  the  pads  will  generally  establish 
the  thickness  of  the  equalizing  lever ;  this  thickness  rarely 
exceeds  l£  inches ;  and  for  small  engines,  say  with  cylin- 
ders 9  inches  in  diameter,  will  sometimes  be  £  inch. 

392.  Before  we  can  determine  the  depth  of  the  equal- 
izing lever  we  must  know  the  load  to  which  it  is  sub- 
jected. Take,  for  instance,  the  spring  over  the  journal 
F  in  Fig.  599.  The  load  on  this  spring  will  be  equal  to 
the  pressure  on  the  journal  F;  one-half  of  this  load  will 
be  transferred  to  the  spring  hanger  A,  the  other  half  to 
the  spring  hanger  B.  In  order  to  produce  equilibrium,  the 
force  acting  at  the  other  end  _&>  °f  the  lever  must  be  equal 
to  that  at  B.  The  force  at  B2  is  equal  to  one-half  of  the 


MODERN  LOCO.VOTITE  CONSTRUCTION.  403 

reaction  of  the  spring  over  the  journal  Fz,  and  this  reaction  is  equal  to  the  pressure 
on  this  journal.  This  mode  of  reasoning  shows  that  the  pressure  on  the  journals  will 
be  equalized ;  in  other  words,  the  pressure  on  the  journal  F  will  be  equal  to  that  on  the 
journal  F.,.  It  further  shows  that  the  tension  on  each  spring  hanger  is  equal  to  one- 
half  of  the  pressure  on  one  joifhial,  or  one-quarter  of  the  sum  of  the  pressures  on  the 
two  journals  F  and  F2 ;  and  the  tension  on  the  fulcrum  D  will  be  equal  to  one-half  the 
sum  of  the  pressures  on  the  two  journals.  Consequently  the  equalizing  lever  E  must 
be  considered  as  a  beam  loaded  at  the  center,  and  freely  supported  at  the  ends,  the 
length  of  the  lever  being  equal  to  the  distance  between  the  centers  of  the  spring 
hangers  B  and  B.2,  and  the  load  equal  to  one-half  the  sum  of  the  pressures  on  the 
journals  F  and  F2.  Thus,  for  instance,  if  the  pressure  on  each  journal  is  1,000 
pounds,  then  the  sum  of  the  pressures  will  be  2,000  pounds ;  the  tension  on  each 
spring  hanger  will  be  500  pounds ;  the  tension  on  the  fulcrum  D  will  be  1,000  pounds, 
and  consequently  the  load  on  the  equalizing  lever  will  also  be  1,000  pounds. 

393.  Spring  gears,  such  as  are  shown  in  Figs.  599  and  606,  cannot  be  used  in 
many  narrow-gauge  engines,  because  in  these  engines  the  width  of  the  fire-box  is 
generally  made  so  great  as  to  leave  insufficient  room  for  the  springs  between  the  fire- 
box and  the  driving  wheels.  Figs.  607,  608  show  a  portion  of  a  general  plan  of  an 
eight-wheeled  passenger  engine,  3  feet  gauge ;  the  spring  gear  is  plainly  shown,  and 
besides  this,  the  design  and  arrangement  of  other  details  are  given,  which  we  shall 
first  describe  very  briefly,  as  we  believe  they  will  be  interesting  and  instructive  to  the 
reader.  The  ash-pan  A  here  shown  is  made  of  sheet-iron.  In  order  to  obtain  a 
sufficient  grate  surface,  the  width  of  the  fire-box  B  is  such  as  tb  cause  the  reduction 
in  the  width  of  those  portions  of  the  frames  F  which  lie  alongside  the  fire-box,  the 
lower  brace  of  the  frame  remaining  the  full  width  throughout.  This  class  of  frames 
we  have  called  "  slab  frames  " ;  they  have  been  illustrated  in  Figs.  298  and  299.  C  is 
the  rear  damper  handle ;  E,  the  rigging  for  the  front  damper  handle ;  G,  the  pads 
securing  the  frames  to  the  boiler,  and  yet  allowing  the  boiler  freedom  for  expansion ; 
//,  the  foot  plate ;  7,  the  injector ;  J,  the  draw  bar ;  K,  the  lever  for  shaking  the  grates ; 
and  L,  the  reversing  lever. 

For  the  spring  gear  two  equalizing  levers  M±  M.2  are  placed  on  the  top  of  each 
driving-box  0;  the  equalizing  levers  M.2  on  the  front  box  are  connected  to  the  spring 
hanger  /'- ;  this  spring  hanger  is  of  the  T  form,  its  upper  branches  forming  hubs  for 
taking  up  the  space  between  the  two  equalizing  levers,  the  whole  being  connected  by  a  pin 
about  £  inch  in  diameter ;  the  lower  end  of  the  hanger  P5  passes  through  slots  cut  in 
the  braces  of  the  frame,  a  spiral  spring  is  placed  between  the  bottom  of  the  frame, 
and  the  gib  which  passes  through  the  lower  end  of  the  hanger.  The  other  ends  of  the 
same  equalizing  levers  are  connected  to  tin-  sin-ing  hanger  P4.  The  form  of  this 
hanger  is  shown  in  Fig.  609.  The  jaws  a  a  are  connected  to  the  levers,  and  the  lower 
end  takes  hold  of  the  spring.  The  other  end  of  the  spring  is  connected  by  means  of  the 
hanger  Pl  to  the  equalizing  lever  A7" below  the  frame;  the  form  of  the  spring  hangers 
P!  and  P:i  is  shown  in  Fig.  610.  The  lever  N  works  on  the  pin  Q  shown  in  Fig.  611 ; 
the  holes  drilled  in  the  boiler  pad  G  and  the  plate  I)  form  bearings  for  the  pin  Q.  A 
pin  or  roller  b  is  inserted  between  the  spring  and  spring  seat  C,  as  shown  in  Fig.  612. 
The  equalizing  levers  on  the  rear  box,  and  spring  hangers  for  the  same,  are  like  those 


MODERN  LOCOMOTITE   COXSTRVCTION. 


405 


we  have  just  described.  All  the  boiler  pads  extend  to  and  are  fastened  to  the  lower 
brace  of  the  frame,  so  as  to  strengthen  the  latter  and  enable  it  to  carry  the  weight 
thrown  upon  it. 

394.  Now  let  us  return  to  the  spring  gear  shown  in  Fig.  599,  and  find  by  calcula- 
tion the  depth  of  the  equalizing  lever  E. 

EXAMPLE  130. — Find  the  depth  of  an  equalizing  lever  for  a  17  x  24  inch  eight- 


n 

in 


Fiy.  612 


Ituilrr 


Fig.  CO9 


Fig.  610 


wheeled  passenger  locomotive;  the  thickness  of  the  lever  is  to  be  l£  inches;   the 
distance  from  the  center  of  the  spring  hanger  B  to  the  center  B2 18  60  inches. 

In  Table  5  (Art.  24)  we  find  that  the  weight  on  the  drivers  for  this  class  and  size  of 
engine  is  52,020  pounds.  From  this  weight  we  must  subtract  the  weight  supported 
directly  by  the  track,  such  as  driving  wheels,  axles,  etc.,  as  stated  in  Art.  391.  Sup- 
posing the  total  weight  to  be  subtracted  (that  of  four  driving  wheels,  two  axles,  four 
driving  boxes,  etc.)  amounts  to  12,000  pounds,  we  then  have  52,020  -  12,000  =  40,020 
pounds  to  be  supported  by  four  springs ;  hence  each  spring  will  have  to  support 


40020 


=  10005  pounds. 


Since  the  load  which  the  equalizing  lever  has  to  support  is  equal  to  one-half  of  the 
sum  of  the  loads  on  two  springs,  we  may  say  that  the  total  load  on  the  lever  is 
10,005  pounds ;  and  since  the  length  of  the  equalizing  lever  is  measured  from  center  to 
center  of  the  spring  hangers,  we  may  consider  the  equalizing  lever  to  be  a  simple  beam 
60  inches  long,  having  a  rectangular  cross-section,  and  freely  supported  at  the  ends ; 
its  thickness  is  l£  inches  and  loaded  at  the  center  with  a  weight  of  10,005  pounds. 
It  is  now  required  to  find  the  depth  of  this  beam.  In  books  treating  on  the  strength  of 
material  we  find  the  following  equation  for  a  beam  of  this  kind,  loaded  under  the 
foregoing  conditions:  J  Wl  =  <jfb  d2,  in  which  TV  represents  the  load  in  pounds;  I, 
the  length  of  the  lever  in  inches;  f,  the  stress  in  pounds  per  square  inch  on  the  outer 
fibers  of  the  lever — that  is,  the  fibers  running  lengthways  of  the  lever ;  6,  the  breadth 
of  the  lever  in  inches ;  and  rf,  the  depth  of  the  lever  in  inches  at  the  center.  Usually 


406  MODERN  LOCOMOTIVE   CONSTRUCTION. 

for  fa  value  of  12,000  pounds  is  assigned,  so  that  OUT-  equation  will  read:  £  x  W  x  I  = 
•g-  x  12000  x  b  x  d2  ;  this  equation  may  be  reduced  to 

W  x  I  =  8000  x  b  x  d2. 
Hence,  to  find  d2  we  have 


8000  x  b 

substituting  the  numerical  values  for  the  symbols,  we  obtain 

10005  x  60 


which,  as  will  be  seen  by  the  symbol  rf2,  is  the  square  of  the  depth  of  the  lever  at  the 
center.     Consequently,  to  find  the  depth  d  we  extract  the  square  root  of  50.02,  or 


d  =  V50.02  =  7.07  inches, 

which  is  the  depth  at  the  center  of  the  equalizing  lever. 

From  the  foregoing  we  may  establish  the  following : 

EULE  98. — To  find  the  depth  at  the  center  of  an  equalizing  lever  for  an  eight- 
wheeled  passenger  engine.  Subtract  from  the  weight  in  pounds  on  the  driving  wheels 
that  portion  of  the  weight  which  is  directly  supported  by  the  rails;  divide  the 
remainder  by  4 ;  the  quotient  will  be  the  load  at  the  center  of  the  equalizing  lever. 
Multiply  this  load  by  the  length  of  the  lever  in  inches,  and  call  the  result  product  A. 
Multiply  the  breadth  in  inches  by  8,000,  and  call  the  result  product  B,  Divide  the 
product  A  by  product  B,  and  find  the  square  root  of  the  quotient ;  the  result  will  be 
the  depth  at  the  center  of  the  equalizing  lever. 

If  there  is  a  notch  cut  across  the  lever,  so  as  to  form  a  bearing  for  the  gib  in  the 
fulcrum,  as  shown  in  Fig.  601,  the  depth  found  by  the  foregoing  rule  will  be  the 
distance  from  the  bottom  of  the  notch  to  the  bottom  of  the  lever. 

If  the  lever  has  a  uniform  breadth,  and  tapers  uniformly  towards  its  ends,  as 
shown  in  Fig.  601,  the  depths  at  the  ends  should  be  equal  to  one-half  of  that  at  the 
center.  This  proportion  will  give  us  an  equalizing  lever  which  is  at  no  place  weaker 
than  at  the  center. 

It  would  seem  that  the  thickness  of  the  metal  at  the  sides  of  the  slots  cut  through 
the  lever  should  be  equal  to  one-half  of  the  thickness  of  the  solid  part  of  the  lever ; 
but  practice  seems  to  indicate  that  better  results  are  obtained  by  increasing  it  about 
15  per  cent. 

The  lever  shown  in  Figs.  601  and  602  was  made  for  a  17  x  24  inch  eight-wheeled 
engine.  Comparing  the  depth  given  in  these  figures  with  the  results  of  our  calcula- 
tion, we  find  that  the  latter  call  for  a  greater  depth.  The  reason  for  this  difference 
is  twofold:  first,  the  weight  on  the  drivers  under  the  engine,  for  which  the  lever  in 
Figs.  601,  602  was  designed,  was  less  than  the  weight  taken  in  o\n-  example ;  second, 
the  lever  must  have  been  made  of  a  better  quality  of  wrought-iron  than  is  usually 
adopted  for  this  class  of  work. 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


407 


408  MODERN  LOCOMOTIVE   CONSTRUCTION. 

If  for  any  reason  the  depth  of  the  equalizing  lever  is  limited,  compelling  us  to 
find  the  thickness  instead  of  the  depth  of  the  lever,  we  may  use  the  following  formula : 

W  x  I 
8000  x  d'2  =    '  *n  w^c^  tt16  Sym1t)ols  denote  the  same  quantities  as  given  on  page  405. 

The  depth  of  the  short  equalizing  lever  in  Fig.  606  is  also  found  by  Eule  98.  In 
regard  to 'the  load  which  the  short  lever  has  to  support  at  its  center,  we  only  need 
to  say  that  it  is  equal  to  the  load  at  the  center  of  the  long  equalizing  lever.  This  will 
be  evident  from  the  fact  that  they  are  connected  by  a  spring. 

395.  Figs.  613,  614  show  a  portion  of  the  sectional  elevation  and  plan  of  a  narrow- 
gauge  Mogul  locomotive.     It  will   be  noticed  that  there  is  not   an  equalizing  lever 
between  the  front  and  central  driving-wheel  springs.     This  arrangement  divides  the 
spring  gear  into  two  distinct  parts.     The  rear  part  is  similar  to  that  described  in  the 
foi-egoing  article ;  the  front  part  consists  of  a  system  of  equalizing  levers  connecting 
the  front  driving-wheel  springs  to  the  two- wheeled  or  pony-truck  springs.     The  front 
spring  hangers  P7  (Fig.  613)  connect  the  springs  to  the  transverse  equalizing  lever  D ; 
the  longitudinal  equalizing  lever  A  is  placed  midway  between  the  frames ;  it  works 
on  the  fulcrum  pin  B  held  in  the  casting  C  bolted  to  the  under-side  of  the  cylinder 
saddle ;  the  rear  end  of  the  lever  A  is  connected  to  the  center  of  the  transverse  lever 
D  by  means  of  the  link  E ;  the  front  end  of  the  lever  A  takes  hold  of  the  king  bolt  in 
the  pony  truck,  which  is  not  shown  here. 

In  wide-gauge  Mogul  engines  the  design  of  the  front  part  of  the  spring  gear  is 
similar  to  the  one  here  shown ;  but  at  the  rear  end,  the  springs  are  generally  placed 
above  the  frames  over  the  driving-wheel  boxes,  and  connected  by  equalizing  levers  as 
shown  in  Fig.  599,  making  the  design  of  the  rear  part  of  the  spring  gear  similar  in  all 
respects  to  that  of  a  driving-wheel  spring  gear  for  a  four-wheeled  passenger  engine. 

In  consolidation  engines  the  springs  over  the  second,  third,  and  fourth  driving 
axle  are  generally  connected  by  equalizing  levers,  forming  a  system  similar  to  the 
design  of  the  driving-wheel  spring  gear  in  a  ten- wheeled  engine,  as  shown  in  Fig.  606. 
The  front  driving-wheel  springs  are  connected  to  the  pony  truck  by  a  system  of  levers 
like  those  used  for  a  Mogul  engine. 

In  eight-wheeled  passenger  engines,  and  Mogul  engines  having  spring  gears  as 
here  described,  the  mass  of  machinery  supported  by  all  the  springs  has  three  main 
points  of  support ;  two  of  these  points  are  at  the  rear  end,  and  the  other  at  the  front. 
In  both  classes  of  engines  the  rear  points  of  support  are  the  fulcrums  of  the  main 
equalizing  levers  midway  between  the  two  rear  driving  wheels.  In  passenger  engines 
the  truck  center  pin  forms  the  front  point  of  support ;  and  in  Mogul  engines  the 
fulcrum  of  the  equalizing  lever  which  connects  the  front  springs  and  the  pony  truck 
forms  the  front  point  of  support.  Ten-wheeled  engines,  and  consolidation  engines 
with  spring  gears  as  here  described,  have  five  main  points  of  support ;  the  fulcrums 
of  the  equalizing  levers  between  the  driving  wheels  form  the  four  main  rear  points  of 
support.  In  a  ten-wheeled  engine  the  front  point  of  support  is  formed  by  the  truck 
center  pin ;  and  in  the  consolidation  engine  the  fulcrum  of  the  equalizing  lever  con- 
necting the  front  driving-wheel  springs  and  the  pony  truck  performs  the  same  office. 

396.  The  center  of  gravity  of  a  passenger  engine  will  always  be  between  the  three 


MODERN  LOCOMOTIVE   CONSTRUCTION.  409 

points  of  support,  as  it  should  be,  in  order  to  make  the  engine  ride  steadily.  The 
center  of  gravity  of  ten-wheeled,  Mogul,  and  consolidation  engines  should  always  lie 
between  the  front  point  of  support  and  the  two  points — one  on  each  side  of  the  engine 
—lying  midway  between  those  driving  wheel  springs  which  are  connected  by  equaliz- 
ing levers.  There  is  danger  of  changing  this  condition;  for  instance,  if  in  Mogul 
engines  all  the  driving-wheel  springs  are  connected  by  equalizing  levers,  the  two  rear 
points  of  support  just  referred  to  will  be  moved  forward  and  the  center  of  gravity 
will  be  at  the  back  of  these  points,  instead  of  lying  between  these  and  the  front 
point  of  support ;  the  consequence  will  be  that  the  rear  end  of  the  engine  will  have  a 
constant  tendency  to  drag  downwards,  and  interfere  with  the  steady  riding  of  the 
engine.  The  same  remarks  apply  to  consolidation  engines.  Here,  then,  we  perceive 
the  reason  for  not  having  equalizing  levers  between  the  front  springs  and  the  adjacent 
ones.  There  is  another  advantage  gained  with  this  arrangement.  When  the  pony- 
truck  wheels  pass  over  any  unevenness  of  the  track,  or  obstacles,  the  truck  springs 
will  be  relieved  of  some  of  the  strain,  a  portion  of  it  being  thrown  on  the  front  driving- 
wheel  springs. 

EQUALIZING-LEVER  FULCRUM. 

397.  The  design  of  the  engine  generally  determines  the  height  of  the  equalizing- 
lever  fulcrum,  which  in  the  majority  of  cases  will  have  to  be  made  as  short  as  possible, 
so  that  the  bulge  of  the  fire-box  will  not  prevent  the  equalizing  lever  from  being  taken 
off  without  removing  the  fulcrum.      This  fact  must  not  be  overlooked  when  the 
distance  from  the  top  of  frame  to  the  center  of  the  boiler  is  to  be  determined;  it 
will  sometimes  compel  us  to  place  the  boiler  higher  than  we  otherwise  should. 
Since   it  is  always  the  aim  to  keep  the  boiler  as  low  as  possible,  it  follows  that 
there  is   no  room  to  spare  for  the  equalizing-lever  fulcrum,  and  its  length  is  cut 
down  so  as  to  leave  sufficient  room  only  for  the  vibration  of  the  equalizing  lever 
without  striking  the  frame  or  top  of 

bolts  which  secure  the  fulcrum  to  the 
frame. 

398.  The  principal  stress  to  which 
the  fulcrum  is  subjected  is  a  tensile 
stress ;  and  since  its  weakest  section  is 
through  a  5,  Fig.  615a,  it  follows  that 
the  area  of  this  cross-section  must  bo 
sufficiently  large  to  resist  the  tensile 

stress.     Careful  observations  indicate  ' ; ; ' 

that  the  general  practice  is  to  allow 

3,000  pounds  per  square  inch  of  cross-section.  The  tensile  stress  on  fulcrum  is  equal 
to  the  load  on  the  equalizing  lever,  consequently,  with  this  data,  the  area  of  the  cross- 
section  through  a  b  (Fig.  615a)  of  the  fulcrum  is  easily  computed  by  the  following 
rule : 

RULE  99. — Divide  the  load  on  the  equalizing  lever  by  3,000 ;  the  quotient  will  be 
the  number  of  square  inches  in  the  cross-section  a  I  of  the  fulcrum. 

EXAMPLE  131. — In  Art.  394  we  have  found  that  the  load  on  the  equiilixing  lever 


1 


410 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


is  10,005  pounds, 
fulcrum. 

Here  we  have 


Find   the  cross-sectional  area  through  the  weakest  part  of  the 


10005 

OAAA  =  3.33+  square  inches,  say  3f  square  inches. 

The  ratio  of  the  thickness  and  width  of  the  fulcrum  is  arbitrary,  and  in  deciding 
upon  these  dimensions  good  judgment  must  be  used ;  but  usually,  on  account  of  the 
cramped  space,  the  fulcrum  will  have  to  be  made  as  thin  as  possible.  If,  now,  the 
thickness  of  the  fulcrum  is  to  be  1  inch,  which  is  about  the  average  thickness,  then 


Fig.  615  c. 

its  width  will  have  to  be  3f  inches  plus  the  thickness  of  the  gib ;  if  the  thickness  of 
the  gib  is  to  be  l£  inches,  then  the  width  of  the  fulcrum  will  have  to  be  3f  +  1|  =  4J 
inches. 

SPUING  HANGEES. 

399.  The  weakest  part  of  a  spring  hanger,  shown  in  Fig.  615&,  is  through  the 
section  e  f.  The  stress  per  square  inch  of  this  area  is  also  3,000  pounds,  the  same  as 
allowed  for  the  fulcrum ;  hence  the  following  rule : 

RULE  100. — Divide  one-half  the  load  on  the  spring  by  3,000 ;  the  quotient  is  the 
number  of  square  inches  in  the  cross-section  through  the  weakest  part  of  the  spring 
hanger. 

EXAMPLE  132. — The  load  which  a  spring  has  to  sustain  is  10,005  pounds.  Find  the 
area  of  the  weakest  part  of  the  spring  hanger. 

Here  we  have 

10005 

— ^—  =  5002.5  pounds  stress  on  the  hanger ; 

and 


5002.5 
3000 


=  1.667  square  inches  in  the  cross-sectional  area. 


MODERX  LOCOMOTIVE  CONSTRUCTION. 


411 


The  ratio  of  its  thickness  and  width  is  also  arbitrary.  If  the  thickness  is 
established,  we  divide  the  cross-sectioual  area  by  the  thickness  ;  the  quotient  plus  the 
thickness  of  the  gib  will  be  the  width  of  the  hanger.  Hence,  if  the  thickness  of  the 
hanger  is  J  inch,  and  the  thickness  of  the  gib  is  ij  inch,  then  the  width  of  the  hanger 
will  be 

'        +  .75  =  2.65  inches,  say  2§  inches. 


If  gibs  are  not  used  as  shown  in  Fig.  615c,  and  if  the  load  on  the  spring  and  the 
thickness  of  the  hanger  remain  the  same  as  before,  then  the  width  of  the  hanger  will 
be  simply 

1.667 
Qr.r   =  L9j  say  2  inches. 

The  foregoing  results  agree  well  with  the  ordinary  practice,  but  it  seems  that 
the  tendency  is  to  make  the  spring  hangers  rather  heavy,  and  the  equalizing  levers 
are  often  made  too  light. 

400.  In  Figs.  603  and  604  are  shown   spring  hangers  made   of  flat  bars;   in- 
deed, this  is  the  simplest  form  of  hangers.     There  are  other  forms  used,  as  shown  in 
Figs.  6156  and  615c.     It  will  be  seen  that  the  one  shown  in  Fig.  615c  has  the  form  of 
an  I,  as  shown  at  B.    When  this  hanger  is  used,  slots  in  the  springs  have  to  be  cut 
through  to  the  ends  as  shown  at  A  ;  and  the  ends  of  the  equalizing  lever  are  forked 
as  shown  at  C.    The  small  projections  forged  to  the  ends  of  the  spring  and  to  the 
ends  of  the  equalizing  lever  prevent  the  hanger  from  disengaging  itself. 

In  Fig.  615ft  the  spring  hangers  have  a  T-lead  forged  to  one  of  their  ends  ;  the 
opposite  ends  are  plain,  and  secured  by  either  pins  or  gibs.  It  will  also  be  noticed 
that  the  pin  through  the  bottom  of  the  hanger  E  does  not  bear 
against  the  under-side  of  the  frame.  In  this  case  two  castings 
a  and  6  are  introduced  between  the  pin  and  the  frame,  with  a 
rubber  block  c,  about  2  to  2£  inches  thick,  between  these  cast- 
ings. Sometimes  a  helical  spring  is  used  in  place  of  the  rubber,  as 
shown  at  P  in  Fig.  607.  This  arrangement  is  somewhat  expen- 
sive, and  is  not  always  used  when  the  main-springs  are  placed 
over  the  axle-box  ;  but  when  the  springs  are  placed  as  shown 
in  Fig.  607,  the  helical  spring  on  the  hanger  must  be  used. 

Sometimes  the  spring  hangers  are  secured  at  their  ends  by 
a  pin  D,  of  the  form  shown  in  Fig.  615d  This  pin  is  cut  from  a 
hexagonal  bar  ;  the  corners  are  turned  off  at  the  center  so  as  to 
form  a  bearing  for  the  hanger  ;  the  corners  which  remain  at  each 
end  form,  so  to  speak,  shoulders  which  prevent  the  pin  from 
moving  out  of  position.  Of  course,  for  a  pin  of  this  kind,  no 
slots  in  the  hangers  are  needed  ;  they  have  simply  holes  drilled 
through  them  large  enough  to  admit  the  hexagonal  part. 

401.  The  gib  through  the  equalizing  fulcrum  tends  to  cut 

into  the  top  of  the  lever  and  thereby  weaken  it.     Consequently  we  often  find  a  steel 
plate  driven  into  the  upper  side  of  the  lever,  as  shown  at  F  in  Fig.  615a.    Another 


to 

o 


9W5- 

_6 

4 

c 

J 

5 

:.!'- 

J 

ep 

^                i     /        ^^^                       ^  ,  

j*                                      ic^«l                                   TM 

-^r 

414 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


mode  of  preventing  the  evil  effects  of  the  wear  under  the  gib  is  to  place  a  loose  steel 
plate  between  the  top  of  equalizing  lever  and  the  gib. 

SPRING  GEAR   USED  BY  THE  PENNSYLVANIA  RAILROAD. 

402.  Fig.  6l5e  shows  a  spring  gear  used  by  the  Pennsylvania  Eailroad  under 
their  standard  eight-wheeled  passenger  engine ;  cylinders,  18£  x  24  inches ;  weight  of 
engine  in  working  order,  106,500  pounds;  weight  on  first  pair  of  drivers,  36,500 
pounds;  weight  on  second  pair  of  drivers,  36,850  pounds;  weight  on  truck,  33,150 
pounds;  steam  pressure,  160  pounds.  The  central  stirrup  A  is  shown  on  a  larger 
scale  in  Fig.  615/;  the  spring-hanger  stirrup  C,  in  Fig.  615/» ;  the  equalizing  fulcrum 
J?,  in  Fig.  615# ;  spring  hanger  Z>,  in  Fig.  615i ;  and  equalizing  lever,  in  Fig.  616. 


rt— H 


*/B *'»  i    r\ 

in r~S" — fn 

**•* H4>< i-*Jf 


jsn 

I 


Fig.  622 


The  gibs  are  shown  in  Fig.  617 ;  the  transverse  brace  for  connecting  the  equalizing 
lever  fulcrum  is  shown  in  Fig.  619;  and  the  cellar  key,  in  Fig.  618.  The  driving- 
wheel  spring  is  shown  separately  in  Fig.  620 ;  several  views  of  the  driving  axle-box 
are  given  in  Fig.  621 ;  and  the  cellar  is  shown  in  Fig.  622. 

The  drawings  show  the  general  arrangement  and  details  so  plainly  that  further 
explanation  is  not  necessary. 


DRIVING-WHEEL   SPRINGS. 

403.  When  the  driving-wheel  springs  are  loaded,  their  lengths  vary  from  30  to 
48   inches;    the    usual    length    is   36    inches.     The   term    "length   of    spring"   is 
always  understood  to  mean  the  distance  a,  Fig.  622a,  from  center  to  center  of  spring 
hangers.     Sometimes  the  term  "  span  "  is  used  in  place  of  length.     The  length  of  spring 
will  depend  much  on  the  design  of  engine ;  no  regular  rule  for  determining  it  can  be 
given,  but  good  judgment  guided  by  experience  must  be  used. 

404.  Springs  36  inches  long  are  generally  3 £  inches  wide;  springs  shorter  than 
36  inches  are  often  made  3  inches  wide ;  and  the  longer  springs,  5  inches  wide,  as  will 
be  seen  in  Fig.  620. 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


415 


405.  The  set  of  spring  is  the  distance  b,  Fig.  622a,  measured  from  the  top,  at  the 
center,  of  the  long  plate  or  leaf  to  the  bearings  of  the  gibs.     The  set  of  the  spring 
when  loaded  should  be  such  that  when  the  axle-box  bears  against  the  top  of  pedestal 
the  leaves  will  be  straight,  as  shown  in  Fig.  6226 ;  and  since  the  play  between  the 
axle-box  and  top  of  pedestal  varies, 

in  different  classes  of  engines,  from 
2i  to  3  inches,  it  follows  that  the 
set  of  springs  under  engines  in 
working  order  will  vary  from  2J  to 
3  inches,  according  to  size  of  en- 
gine. 

406.  The  usual  thickness  of  the 
leaves  is  |  inch,   sometimes  it  is 
£  inch;    the   former   thickness  is 
often  adopted  for  springs  36  inches 
long ;  for  springs  much  longer  than 
this,   the  leaves  are  made  J  inch 
thick ;  and  for  comparatively  short 

springs  the  thickness  is  -fa  inch.  The  greater  the  thickness  the  greater  will  be  the 
stress  on  the  outer  fibers  of  the  plate ;  hence,  for  a  durable  spring  the  thickness  of 
plates  should  be  reduced  as  much  as  is  consistent  with  good  practice. 

407.  After  the  length,  width  of  spring,  and  thickness  of  leaves  have  been  decided 
upon,  the  number  of  leaves  will  have  to  be  computed ;  this  can  be  done  by  the  follow- 
ing rule : 

RULE  101. — Multiply  the  load  in  tons  (2,000  pounds  per  ton)  which  the  spring  has 
to  sustain  by  the  length  of  spring  in  inches,  and  multiply  this  product  by  11 ;  call  this 
product  A.  Multiply  the  width  of  the  spring  in  inches  by  the  square  of  the  thickness 
in  sixteenths  of  an  inch  of  one  leaf ;  call  this  product  B.  Divide  product  A  by  the  prod- 
uct B ;  the  quotient  will  give  the  number  of  leaves  required.  Or,  in  symbols,  we  have 

Load  in  tons  x  length  of  spring  in  inches  x  11 
Breadth  in  inches  x  "(thickness  of  one  leaf  hTsixteenths)2  = 

The  load  on  the  spring  is  found  in  the  same  manner  as  the  load  was  found  on  the 

_y— _^_^     equalizing  lever,  explained  in  Art. 
— p^-*—1'        394;  in  fact,  the  load  on  the  spring 
is  equal  to  the  load  on  the  equal- 


Fit'. 


izing  lever. 

EXAMPLE  133.— The  load  which 
a  spring  has  to  sustain  is  5  tons ;  length  of  spring,  36  inches ;  width,  3J  inches ;  thick- 
ness of  leaves,  g  inch.     Find  the  number  of  leaves  required  to  sustain  the  given  load. 
According  to  the  foregoing  rule,  we  have 

5  x  36  x  11 
3.5  x  62          15'7' 

This  answer  calls  for  more  than  15  leaves,  and  less  than  16 ;  to  be  on  the  safe  side, 
it  is  advisable  to  adopt  16  leaves. 


41G  MODERN  LOCOMOTIVE   CONSTRUCTION. 

408.  In  ordering  a  spring  it  is  necessary  to  know  its  deflection,  so  that  the  set 
without  the  load  can  be  given.  The  following  rule  may  be  used  for  finding  the 
deflection : 

EULE  102. — Multiply  the  cube  of  the  length  in  inches  by  1.5 ;  call  this  product  A. 
Multiply  the  width  of  the  plate  in  inches  by  the  cube  of  its  thickness  in  sixteenths, 
and  multiply  this  product  by  the  number  of  plates  or  leaves ;  call  this  product  B. 
Divide  product  A  by  product  V ;  the  quotient  will  be  the  deflection  in  sixteenths  of  an 
inch  per  ton  of  2,000  pounds. 

In  symbols,  we  have 

(Length  of  spring  in  inches)'   x  1.5 

=  deflection  in  sixteenths  per  ton. 

Width  of  plate  in  inches  x  (thickness  in  sixteenths)''   x  number  of  plates 

EXAMPLE  134. — The  length  of  spring  is  36  inches ;  width,  3£  inches ;  thickness  of 
plates,  f  inch ;  number  of  plates,  16 ;  load  to  be  sustained,  5  tons.  Find  the  deflection 
of  the  springs. 

Here  we  have 

363  x  1 5 
o~-~~;  ~K3 — ~T«  =  5-79  sixteenths  of  an  inch  deflection  per  ton. 

o.O    X    O'    X    J.O 

The  load  is  5  tons,  hence  the  total  deflection  is  5.79  x  5  =  28.9  sixteenths,  say  ft  = 
IT!  inches  deflection  for  a  load  of  5  tons.  According  to  Art.  405,  the  spring  should 
have  a  set,  when  loaded,  equal  to  the  play  between  top  of  axle-box  and  top  of 
pedestal ;  if  this  is  equal  to  2£  inches,  then  the  total  set  before  the  spring  is  loaded 
should  be  2£  +  Iff  =  4-fy  inches. 


CHAPTER  X. 

BOILERS.— GRATE     SURFACE.— HEATING     SURFACE— RIVETED    JOINTS.— EXTEN- 
SION  FRONTS. 

BOILERS. 

409.  Figs.  623,  624,  625,  626  represent  different  views  of  a  boiler  for  an  eight- 
wheeled  passenger  engine  with  cylinders  18  inches  diameter  and  24  inches  stroke. 
Figs.  633,  634,  635  represent  a  boiler  and  its  details  for  a  consolidation  engine  with 
cylinders  20  inches  diameter  and  24  inches  stroke.    Both  boilers  are  designed  for 
burning  bituminous  or  soft  coal. 

In  all  boilei-s  provision  must  be  made  for  cleaning  and  washing,  and  for  this 
purpose  every  boiler  must  be  provided  with  hand  holes  at  each  corner  of  the  fire-box ; 
they  are  generally  placed  between  the  fire-box  ring  a,  Fig.  623,  and  the  first  row  stay 
bolts ;  another  hand  hole  must  be  placed  in  the  bottom  of  the  front  tube  sheet,  as 
shown  at  i,  Fig.  626.  Occasionally  we  find  two  or  three  hand  holes  placed  in  each 
side  of  the  fire-box  about  1  inch  above  the  crown  sheet.  We  have  seen  a  brass  plug 
placed  in  the  fire-box  flue  sheet  directly  under  the  flues,  but  this  practice  should  be 
condemned,  because  plugs  in  the  fire-box  are  liable  to  burn  out,  causing  considerable 
trouble  and  annoying  delays. 

410.  The  hand  holes  ai-e  made  either  oval  or  circular  in  form.     The  average  size 
of  oval  holes  is  shown  in  Fig.  632.    The  forms  of  plates  for  these  holes  are  shown  in 
Figs.  641,  642,  643.     They  are  made  of  cast-iron,  with  a  wrought-iron  bolt  through 
their  centers  to  clamp  them  to  the  sheet.     The  average  sizes  of  circular  holes  and  mud 
plugs  used  in  place  of  oval  ones  are  shown  in  Fig.  637.     The  mud  plugs  are  usually 
made  of  brass ;  they  have  generally  a  square  recess,  sometimes  a  square  projection,  for 
screwing  them  into  the  holes.     Plugs  with  recesses  are  to  be  preferred,  because  they 
take  up  less  room.     The  taper  of  the  plugs  is  about  1  inch  in  12  inches ;  the  thread  is 
of  a  comparatively  fine  pitch,  12  or  14  threads  per  inch.     Sometimes  reinforcing  plates 
are  used,  as  shown  in  Fig.  637.     They  are  riveted  to  the  inside  of  the  sheet,  so  as  to 
obtain  more  perfect  threads  in  the  holes,  but  this  is  not  a  common  practice,  although 
it  is  a  good  one. 

411.  A  mud  drum  is  sometimes  riveted  to  the  bottom  of  the  boiler  near  the  front  flue 
sheet,  as  shown  in  Fig.  644.     The  diameter  of  the  drum  is  generally  15  inches ;  depth, 
from  10  to  12  inches.     A  wrought-iron  ring  is  riveted  to  the  outside,  at  the  bottom 
end,  to  which  a  cast-iron  head  is  bolted;  the  bolts  should  have  large  squareheads, 
one  of  their  sides  bearing  against  the  outside  of  the  drum  to  prevent  them  from 


418 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


turning  in  screwing  up  the  nuts,  otherwise  the  lagging  around  the  drum  may  have  to 
be  removed  (which  is  not  desirable)  when  the  drum  head  is  to  be  taken  off.  A  blow- 
off  cock  is  screwed  into  the  center  of  the  head,  and  worked  by  means  of  a  rod  extend- 


O!9       OOOO.OO       O  O 

OOOOO.OO  O 
O      O    '  O      O    0   O  C 

O   O    O  O     D  tj3      O  O      O 


L>- 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


419 


ing  to  the  outside  of  the  frames.  The  writer  does  not  favor  the  use  of  mud  drums, 
and  believes  that  better  results  will  be  obtained  by  screwing  the  blow-off  cock 
directly  into  the  bottom  of  the  boiler,  in  about  the  same  position  as  that  occupied  by 
the  mud  drum.  Although  this  is  not  a  general  practice,  we  believe  it  to  be  a  good 
one,  because  the  check  valves  are  nearly  always  placed  near  the  front  end,  as  indicated 
at  c  in  Fig.  623,  and  the  mud  and  other  impurities  which  are  carried  with  the  water 


Fig.  02  1 


Fig.  632 


Fig!  631 


into  the  boiler  will  settle  directly  under  the  check- valve  openings  and  often  pile  up  in 
a  wedge  form,  highest  at  the  front  flue  sheet,  forming  nearly  an  angle  of  45  degrees 
with  the  same ;  hence  the  utility  of  having  a  blow-off  cock  at  the  front  end  of  the 
boiler  must  be  readily  perceived.  Whether  or  not  a  blow-off  cock  is  placed  there,  one 
is  always  attached  to  the  fire-box  as  low  as  possible ;  it  is  screwed  in  the  back  head 
of  the  boiler  whenever  the  driving  axle  will  permit  its  use  there ;  sometimes  it  is 
placed  in  the  front  of  the  fire-box,  and  occasionally  in  the  sides ;  it  is  placed  in  the 
latter  position,  either  because  there  is  no  room  for  it  anywhere  else,  or  because  in 
some  designs  of  engines  it  is  more  convenient  to  have  it  there. 


H 
«X 


MOVER!?  LOCOMOTIVE  CONSTRUCTION. 


421 


41±  It  may  also  be  well  to  notice  that  the  hand-hole  plates  or  mud  plugs  in  the 
bottom  of  the  fire-box  should  be  placed  iu  such  positions  as  to  make  access  to  them 
easy,  and  convenient  for  washing  and  cleaning  the  water  space  around  the  furnace. 

• — *4  //  " 

_ nil  I .  _       -  I  M 3V- M 


When  the  fire-boxes  extend  downwards  between  the  driving  axles,  it  is  often  difficult 

to  place  the  hand-hole  plates  in  desirable  positions,  particularly  at  the  rear  end, 

because  the  rear  driving  axle  is  generally  closer  to  the  fire-box  than  the  front  axle ; 

the  distance  from  the  center  of  the  rear  axle 

to  the  fire-box  varies  from  6  to  7  inches, 

according  to  the  size  of  engine,  leaving  in 

many  cases  a  clearance  of  only  £  inch  between 

the  flanges  of   the  rear  driving-boxes    and 

fire-box,  thereby  compelling  us  to  be  satisfied 

with  such  positions  of  the  hand  holes  as  are 

not  exactly  desirable. 


GKATE   SUEFACE. 


ISrfi— j 


Fig 


644 


413.  Locomotive  boilers  may  be  divided 
into  two  classes,  namely,  bituminous  or  soft- 
coal  burners,  and  anthracite  or  hard-coal  burners.    We  shall  consider  the  soft-coal 
burners  first. 

In  a  large  number  of  eight-wheeled  passenger,  Mogul,  and  ten-wheeled  engines, 
the  fire-box  is  placed  between  the  two  rear  axles,  and  in  a  few  of  the  larger  engines  it 
is  placed  above  the  axles,  extending  over  the  rear  one.  The  advantage  of  placing  the 
fire  between  the  axles  is  that  a  deep  fire-box  can  be  obtained  so  as  to  accommodate  the 
depth  of  soft-coal  fires ;  the  depth  of  these  fires  varies  from  15  to  24  inches,  according 
to  the  quality  of  the  coal  used  and  the  nature  of  the  service.  If  the  fire-box  is  not 
placed  between  the  axles,  then  in  order  to  obtain  the  required  depth  of  fire  the  boiler 
must  be  raised.  For  fast  running  engines  it  is  always  advisable  to  keep  the  boiler  as 
low  as  possible,  and  since  eight-wheeled  passenger  engines  are  generally  designed  for 
high  speeds,  we  do  not  often  find  the  fire-box  in  this  class  of  engines  placed  above 
the  rear  driving  axle. 

Since  the  available  space  in  passenger  engines,  say  with  cylinders  18  inches 
diameter  and  upwards,  is  insufficient  for  a  size  of  fire-box  which  can  give  the  highest 
economy  of  fuel,  it  follows  that  great  care  must  be  taken  not  to  give  a  greater 


422  MODERN  LOCOMOTIVE   CONSTRUCTION. 

amount  of  clearance  between  the  fire-box  and  other  mechanism  than  is  absolutely 
necessary.  The  distance  between  the  frames  limits  the  width  of  the  fire-box  ;  hence 
in  designing  a  large  eight-wheeled  passenger  engine  which  is  to  have  the  fire-box 
between  the  axles,  all  that  can  be  done  is  to  exercise  good  judgment  in  determining 
the  distance  between  the  driving  axles,  which,  as  we  have  seen  in  Art.  205,  is  limited 
by  the  length  of  the  side-rod,  and  then  use  the  available  space  to  the  best  advantage. 
In  smaller  engines,  and  also  in  locomotives  which  have  the  fire-box  above  the  axles,  the 
space  for  the  fire-box  is  not  restricted  to  such  a  small  extent,  and  in  such  cases  the 
question  arises  :  What  is  the  proper  area  for  the  grate  surface  ? 

414.  It  is  impossible  to  give  any  hard-and-fast  rule  for  computing  the  grate  area 
suitable  for  the  many  various  conditions  of  service  and  character  of  fuel.  But  it  is 
reasonable  to  assume  that  the  grate  surface  should  be  governed  by  the  horse-power 
developed  when  the  engine  is  exerting  its  maximum  continuous  effort.  To  compute 
the  horse-power  we  must  know  the  piston  speed  and  the  mean  effective  pressure  on  the 
piston.  But  the  horse-power  is  also  limited  by  the  tractive  force,  therefore  locomotive 
builders  usually  adopt  a  certain  ratio  of  grate  surface  to  the  tractive  force  ;  but  they  do 
not  agree  on  this  ratio,  and  the  consequence  is  that  it  varies  considerably.  We  find 
that  an  allowance  of  1  square  foot  of  grate  surface  for  every  600  pounds  of  tractive 
force  agrees  very  closely  with  the  average  practice.  Hence  we  can  establish  the  fol- 
lowing rule  : 

EULE  103.  —  Divide  the  tractive  force,  as  found  by  Rule  3,  by  600  ;  the  quotient 
will  be  the  number  of  square  feet  of  grate  area  for  soft-coal  burning  engines. 

EXAMPLE  135.  —  Find  the  number  of  square  feet  of  grate  surface  for  a  soft-coal 
burning  passenger  engine  with  cylinders  15  x  24  inches;  driving  wheels,  55  inches 
diameter  ;  mean  effective  pressure  on  the  piston,  90  pounds  per  square  inch. 

According  to  Eule  3,  the  tractive  force  of  this  engine  is 

152  x  90  x  24 

-  ^p—       -  =  8836  pounds. 
oo 

According  to  Eule  103,  we  have 

8836 

=  14.72  square  feet  for  the  grate  area. 


If,  now,  the  furnace  is  to  be  35  inches  =  2.91  feet  wide,  then  its  length  will  have 
to  be  14  72 

~£9f 

EXAMPLE  136.  —  Find  the  number  of  square  feet  of  grate  area  for  a  soft-coal 
burning  passenger  engine  with  18  x  24  inch  cylinders  ;  drivers,  61  inches  diameter  ; 
mean  effective  steam  pressure  on  the  piston,  90  pounds  per  square  inch. 

The  tractive  force  of  this  engine  is 

182  x  90  x  24 

—  TTJ  —       ~  =  11472  pounds. 


And  the  grate  surface  will  be 

11472 
600 


=  19.12  square  feet. 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


423 


If,  now,  the  width  of  the  furnace  is  to  be  2.91  feet,  then  its  length  will  have  to  be 

19.12 


2.91 


=  6.57  feet. 


If  there  is  no  room  for  this  length  of  furnace  between  the  axles,  then  we  must  place 
tli«'  furnace  above  the  axles,  or  be  satisfied  with  less  grate  surface. 

In  general,  if  the  grate  surface  determined  by  the  foregoing  rule  is  too  large  for  a 
fire-box  to  be  placed  between  the  axles,  then  we  must  place  it  above  them,  or  if  this 
cannot  be  done,  we  must  be  either  satisfied  with  a  reduced  steaming  capacity  of  the 
boiler,  or  reduce  the  area  of  the  exhaust  nozzles,  which  will  to  some  extent  make  up 
for  the  loss  sustained  by  an  insufficient  grate  area.  Although  this  mode  of  procedure 
is  a  general  one,  it  is  by  no  means  free  from  objections,  because  the  reduction  of  the 
diameter  of  the  exhaust  nozzle  not  only  increases  the  back  pressure  in  the  cylinders, 
but  it  also  tends  to  lift  the  fuel,  and  draw  some  of  it  unconsumed  through  the  flues, 
thereby  creating  a  waste  of  fuel.  Even  with  a  grate  area  as  large  as  is  found  by  the 
given  rule  we  shall  need  a  blast  of  considerable  force. 

The  grate  area  in  the  following  tables  have  been  computed  by  Eule  103;  the 
mean  effective  steam  pressure  on  the  piston  has  been  taken  at  90  pounds  per  square 
inch  for  all  the  engines.  It  will  be  found  that  these  grate  areas  agree  very  closely 

with  the  average  practice. 

TABLE  59. 

CALCULATED  GRATE  AREA   FOR  SOFT-COAL   BURNING  EIGHT-WHEELED  PASSENGER  ENGINES. 


Size  of  Cylinders. 

Diameter  of  Driving  Wheels. 

Grate  Area. 

Column  1. 

Column  i. 

Column  3. 

Inches. 

Inches. 

Square  feet. 

10  x  20 

45 

6.6 

11  x  22 

45 

8.8 

12  x  22 

48 

9.9 

13  x  22 

49 

11.3 

14  x  24 

55 

12.8 

15  x  24 

55 

14.7 

16  x  24 

58 

15.8 

17  x  24 

60 

17.3 

18  x  24 

61 

19.1 

TABLE   60. 

CALCULATED  GRATE  AREA  FOR  SOFT-COAL  BURNING  CONSOLIDATION  ENGINES. 


Size  of  Cylinder-. 

Diameter  of  Driving  Wheels. 

Grate  Area. 

Column  1. 

Column  2. 

Column  3. 

Inches. 

Inches. 

Square  feet. 

14  x  16 

36 

13.0 

15  x  18 

36 

16.8 

20  x  24 

48 

30.0 

22  x  24 

50 

34.8 

Comparing  the  grate  area  of  the  engines  with  cylinders  14  and  15  inches  diameter, 
given  in  Table  60,  with  the  grate  area  of  engines  having  cylinders  14  and  15  inches 


424 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


in  diameter,  as  given  in  Table  59,  we  find  that  the  length  of  stroke  and  the  diameter 
of  wheels  affect  the  grate  area  to  small  extent  only.  Hence,  a  number  of  locomotive 
builders  adopt  the  same  grate  area  for  a  given  size  of  cylinder  for  all  classes  of  soft- 
coal  burning  engines. 

415.  Usually  when  the  fire-box  is  placed  above  the  axles,  it  is  still  kept  between 
the  frames,  which,  of  course,  limits  its  width  but  gives  us  more  available  space  for 

length.  Sometimes  this  practice  is 
changed  by  placing  the  fire-box  en- 
tirely above  the  top  of  frames;  such 
a  course  will  give  us  a  greater 
width,  but  it  will  also  raise  the  boiler 
higher  than  it  would  be  by  keeping 
the  fire-box  between  the  frames  and 
placing  it  simply  above  the  axles.  To 
overcome  the  objection  to  raising  the 
boiler  too  high,  the  fire-box  has  been 
made  of  a  form  as  shown  in  Figs.  645 
and  646,  which  are  so  plain  that  fur- 
ther explanation  is  unnecessary.  This 
design  was  brought  out  by  Mr.  John 
Headden,  during  the  time  he  held 
the  position  of  superintendent  of  the 
Rogers  Locomotive  Works.  Several  of 
these  locomotives  have  been  built,  and, 
as  far  as  we  know,  have  given  good 
satisfaction ;  yet  this  design,  although 
it  embraces  excellent  points,  is  not 
free  from  objections — for  instance,  the 
top  of  the  rear  end  of  frames  must 
be  made  somewhat  crooked ;  and  sec- 


Fig. '646 


ondly,  the  springs  cannot  be  placed  above  the  frames,  which,  after  all,  seems  to  be  the 
favorite  arrangement. 

ANTHEACITE,   OE  HAED-COAL  BURNING  ENGINES. 

416.  In  this  class  of  engines  we  require  a  larger  grate  surface  than  in  bituminous 
or  soft-coal  burning  engines,  because  in  the  former  shallower  fires  are  carried,  ranging 
from  6  to  15  inches  deep,  according  to  the  quality  of  coal  and  the  nature  of  the  service 
of  the  engine,  while  in  soft-coal  burning  engines  the  depth  of  fires  varies  from  15  to  24 
inches.  Since  both  kinds  of  coal  contain  about  the  same  number  of  units  of  heat, 
in  round  numbers  about  14,000,  it  follows  that  in  order  to  evaporate  the  same  amount 
of  water  with  both  kinds  of  coal,  the  grate  surface  for  the  hard  coal  must  be  greater 
than  that  for  soft  coal. 

The  reason  for  the  difference  in  the  depths  of  fires  may  be  stated  as  follows: 
When  bituminous  coal  is  heated  in  the  furnace  it  parts  readily  with  the  hydrocarbons 
in  the  form  of  gas,  leaving  the  solid  portion  of  coal  as  a  spongy,  porous  mass  (coke) ; 


MODESN  LOCOMOTIVE   CONSTRUCTION. 


425 


r 


426  MODERN  LOCOMOTIVE   CONSTRVCTION. 

the  lumps  break  up  into  smaller  pieces,  allowing  the  air  to  come  in  contact  with  the 
interior  of  the  solid  coal. 

The  anthracite  coal,  on  the  contrary,  does  not  break  to  pieces  so  readily,  if  at  all, 
thereby  permitting  the  air  to  come  in  contact  with  the  exterior  surface  only.  Hence, 
this  coal  requires  a  freer  intermingling  of  air  with  it,  and  therefore  a  comparatively 
shallow  fire  must  be  carried.  But  for  shallow  fires  the  depth  of  furnace  can  be 
reduced,  enabling  us  to  place  the  fire-box  above  the  axles  without  raising  the  boiler  to 
an  undesirable  height.  Hence  in  hard-coal  burners  the  fire-boxes  are  generally  placed 
above  the  axles. 

417.  The  grate  area  adopted  for  this  class  of  engines  by  the  different  builders  also 
varies  considerably.  We  find  that  the  tractive  force  as  found  by  Eule  3,  divided  by 
500,  will  give  us  a  grate  area  in  square  feet  which  agrees  very  closely  with  the  average 
practice.  Hence  we  have  the  following  rule: 

EULE  104.  —  Divide  the  tractive  force  as  found  by  Rule  3,  by  500  ;  the  quotient 
will  be  the  grate  surface  in  square  feet  for  hard-coal  burning  engines. 

EXAMPLE  137.  —  Find  the  number  of  square  feet  of  grate  surface  for  a  hard-coal 
burning  eight-wheeled  passenger  engine  having  cylinders  15  inches  diameter,  24 
inches  stroke  ;  driving  wheels,  55  inches  diameter  ;  mean  effective  pressure,  90  pounds 
per  square  inch  of  piston. 

According  to  Rule  3,  the  tractive  foi'ce  is 

152  x  90  x  24 

-—  -       -  =  8836  pounds, 
oo 

and 

8836 

—  17.6  square  feet  of  grate  surface. 


If  the  width  of  grate  is  to  be  2.91  feet,  then  its  length  will  be 

17.6 


2.91 


=  6  feet. 


EXAMPLE  138. — Find  the  number  of  square  feet  of  grate  area  for  a  hard-coal 
burning  Mogul  engine  with  cylinders  18  inches  diameter,  24  inches  stroke;  driving 
wheels,  51  inches  in  diameter ;  mean  effective  pressure,  90  pounds  per  square  inch  of 
piston. 

The  tractive  force  of  this  engine  is  13722.3  pounds,  hence  the  grate  surface  will  be 

13722.3 

— _AA     =  27.44  square  feet. 

OUU 

If  the  grate  is  to  be  2.91  feet  wide,  its  length  will  be 

27.44 


2.91 


=  9.4  feet. 


EXAMPLE  139. — Find  the  grate  area  for  a  hard-coal  burning  consolidation  engine 
with  cylinders  20  inches  diameter,  24  inches  stroke ;  driving  wheels,  48  inches  diameter ; 
mean  effective  pressure,  90  pounds  per  square  inch  of  piston. 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


427 


The  tractive  force  of  this  engine  is  18,000  pounds,  hence  the  grate  surface  will  be 

18000 


500 


=  36  square  feet. 


If  the  width  of  grate  is  to  be  2.91  feet,  then  its  length  will  be 

36 


2.91 


=  12.3  feet. 


The  grate  areas  in  the  following  tables  have  been  computed  by  the  foregoing  rule. 

TABLE    61. 

CAI-CULATED   GRATE  AREA   FOR   HARD-COAL  BURNING  EIGHT-WHEELED  PASSENGER   ENGINES. 


Size  of  Cylinders. 

Diameter  of  Driving  Wheels. 

Orate  Area. 

Coin  inn  1. 

Column  2. 

Column  3. 

Incbca. 

Inches. 

Square  feet. 

10  x  20 

45 

8.0 

11  x  22 

45 

10.6 

12  x  22 

48 

11.8 

13  x  22 

49 

13.6 

14  x  24 

55 

15.3 

15  x  24 

55 

17.6 

16  x  24 

68 

19.0 

17  x  24 

60 

20.8 

18  x  24 

61 

22.9 

TABLE    62. 

CALCULATED   GRATE  AREA  FOR  HARD-COAL  BURNING   CONSOLIDATION   ENGINES. 


Size  of  Cylinders. 

Diameter  of  Driving  Wheels. 

Orate  Area. 

Column  1. 

Column  2. 

Column  8. 

Inches. 

Inches. 

Square  feet. 

14  x  1C 

36 

15.6 

15  x  18 

36 

20.2 

20  x  24 

48 

36.0 

22  x  24 

50 

41.8 

The  tendency  is  to  make  the  grate  area  for  the  larger  hard-coal  burning  engines  a 
little  less  than  given  in  the  tables,  so  as  not  to  get  a  length  of  furnace  which  is  incon- 
venient for  firing ;  but  for  the  smaller  engines  the  tendency  is  to  make  the  grate  area 
a  little  greater  than  given  in  the  tables,  as  such  furnaces  will  not  be  too  long  to  fire. 
Furnaces  over  12  feet  in  length  are  difficult  to  fire,  and  should  be  avoided. 

GREATEST   WIDTH   OF   FIRE-BOX. 

418.  In  all  ordinary  wide-gauge  (4'  "  8£")  locomotives  the  outside  width  of  the  fire- 
box cannot  exceed  the  distance  from  the  inner  to  outer  sides  of  the  frames,  because 
the  clearance  between  the  tires  and  the  frames  is  usually  from  1  to  1J  inches,  and 


428  MODERN  LOCOMOTIVE    CONSTRUCTION. 

less  clearance  than  this  between  the  fire-box  and  the  tires  is  impracticable ;  even  with 
this  clearance  the  rivets  in  the  fire-box,  in  the  neighborhood  of  the  tires,  will 
frequently  have  to  be  countersunk  to  prevent  contact. 


DEPTH  OF  HARD  AND   SOFT  COAL  FURNACES. 

419.  The  same  practical  considerations  which  determine  the  depth  of  a  soft-coal 
burning  furnace  will  also  determine  the  depth  of  a  hai'd-coal  burning  furnace.  The 
depth  of  the  furnace  will  depend,  first,  on  the  number  of  flues  and  the  depth  of  the 
fires ;  second,  it  will  depend  on  the  clearance  which  must  be  given  between  the  bulged 
part  of  the  fire-box  and  the  springs  or  the  flanges  of  the  wheels ;  third,  it  will  depend 
on  the  manner  of  fastening  the  frames  to  the  fire-box. 

First.  To  find  the  depth  of  the  furnace,  we  must  lay  in  the  tubes  as  shown  in 
Fig.  624,  and  then  lay  in  the  flange  of  the  tube  sheet  so  that  its  inner  surface  will  be 
about  $  of  an  inch,  or  not  more  than  1  inch  above  the  top  of  the  upper  row  of  tubes ; 
in  this  way  we  establish  the  position  of  the  crown  sheet.  .  We  then  add  the  required 
depth  below  the  flues  for  the  fuel,  and  this  generally  establishes  the  necessary  depth  of 
the  furnace.  If,  now,  the  fire-box  is  to  be  placed  on  top  of  frames,  we  have  simply  to 
raise  the  boiler  high  enough  for  this  depth  of  furnace. 

Second.  In  hard-coal  burning  engines  the  boiler  will  generally  set  so  high  as  to 
clear  the  springs  and  the  flanges  of  the  driving  wheels,  but  to  be  certain,  it  is  always 
best  to  lay  in  the  frames  as  shown  in  Fig.  654,  and  also  draw  in  the  springs  and  wheels 
in  the  position  they  will  occupy  when  the  engine  is  in  working  order.  There  should 
be  at  least  4  inches  clearance  between  the  flanges  a  of  the  wheels  and  the  bulge  b,  so  as 
to  allow  for  the  vertical  movement  of  the  wheels  in  the  pedestals ;  there  should  also  be 
about  4  inches  clearance  between  the  ends  of  the  springs  (not  shown  here)  and  the  bulge, 
so  as  to  allow  the  springs  to  be  placed  in  position  before  they  are  strapped  down.  If 
there  is  less  than  4  inches  of  clearance  the  boiler  should  be  raised  bodily  and  the  fire- 
box made  deeper,  allowing  the  bottom  to  come  within  about  1  inch  above  the  top  of 
frames.  It  is  also  necessary  to  lay  in  the  lifting-shaft  and  see  that  there  is  clearance 
between  the  link  motion  and  the  lagging  around  the  barrel  of  the  boiler.  For 
soft-coal  burning  engines  we  proceed  in  pi'ecisely  the  same  way,  and  then  provide  for 
the  third  condition,  which,  as  we  have  seen,  is,  that  the  depth  of  the  furnace  will 
depend- on  the  manner  of  fastening  the  fire-box  to  the  frames.  Of  course,  when  the 
fire-box  is  placed  above  the  frames,  and  we  have  met  all  the  practical  requirements 
just  mentioned,  the  manner  of  fastening  the  fire-box  to  the  frames  will  not  affect  the 
depth  of  the  furnace.  But  when  the  fire-box  is  placed  between  the  frames  as  is  shown 
in  Fig.  654,  it  is  customary  to  allow  its  lower  end  c  to  extend  a  short  distance  below 
the  lower  braces  d  of  the  frames  so  as  to  bring  the  heads  of  the  rivets  through  the  ring 
of  the  fire-box  below  the  under  side  of  the  lower  braces,  otherwise  the  heads  of  the 
rivets  will  in  many  cases  have  to  be  countersunk,  which  in  good  locomotive  practice 
is  avoided  as  much  as  possible.  This  arrangement  allows  the  pads  to  be  attached  to 
the  lower  brace  of  the  frame  as  well  as  to  the  upper  one,  thereby  securing  the  frames 
rigidly  in  a  lateral  position,  the  fire-box  acting  as  a  brace  between  them.  Sometimes 
the  required  depth  of  the  ash-pan /will  not  allow  the  fire-box  to  extend  so  far  down- 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


429 


wards ;  in  such  cases  the  bottom  c  of  the  fire-box  just  reaches  the  top  g  of  the  lower 
brace  of  the  frame. 

The  provisions  made  for  fastening  the  fire-box  to  the  frames  as  just  explained 
may  give  us,  in  some  cases,  a  deeper  furnace  than  is  required  for  the  depth  of  the  fire ; 


noonn 
Sooooc 


ogogogog 

i    on    o 

Oo°o°o0 
gogogo 


it  is  for  this  reason  that  wo  have  Paid  that  the  dopth  of  tho  furnace  will  depend  on  the 
manner  of  fastening  the  fire-box  to  the  fram.-s.  This  class  of  boilers  is  often  used  for 
passenger  engines,  and  since  the  driving  wheels  of  this  class  of  engines  are  of  a  large 
diameter  comparatively,  there  is  sometimes  trouble  in  finding  room  for  the  reach-rod  h 


430  MODERN  LOCOMOTIVE   CONSTRUCTION. 

(Fig.  654),  which  connects  the  reversing  lever  i  to  the  lifting-shaft ;  hence  the  cross- 
section  of  the  reach-rod  should  also  be  drawn  in  the  end  view,  and  sufficient  room 
allowed  for  it ;  to  do  so  may  sometimes  compel  us  to  raise  the  boiler  a  little.  Now, 
having  provided  for  all  these  practical  requirements,  we  have  not  only  established  the 
depth  of  the  furnace,  but  we  have  also  established  the  height  of  the  boiler  above  the 
frames;  for  we  cannot  place  it  lower  without  interfering  with  the  action  of  other 
mechanism,  and  we  cannot  place  it  higher  without  reducing  the  stability  of  the  engine ; 
the  boiler  should  always  be  kept  as  low  as  possible. 

FIKE-BOXES  FOE  WOOD-BURNING  LOCOMOTIVES. 

420.  For  wood-burning  locomotives  a  very  deep  fire-box  is  required ;  consequently, 
in  this  class  of  engines,  it  is  nearly  always  placed  between  the  axles,  thereby  obtaining 
the  same  size  of  grate  surface  as  is  used  for  soft-coal  burning  engines  with  fire-boxes 
placed  in  a  similar  position. 

SLOPING  CEOWN   SHEETS. 

421.  When  a  locomotive  has  to  run  down  a  hill,  the  water  level  in  the  boiler  will 
of  course  remain  horizontal,  hence  the  depth  of  the  water  at  the  rear  end  of  the  engine 
above  the  center  line  of  the  boiler  will  be  less  than  at  the  front  end.     If,  now,  the  crown 
sheet  has  been  placed  horizontally,  that  is  to  say,  parallel  with  the  center  line  of  the 
boiler,  there  is  danger  of  the  rear  end  of  the  crown  sheet  projecting  above  the  water 
level,  which  may  cause  it  to  be  burnt,  and  lead  to  serious  or  fatal  accidents.     In  order 
to  keep  it  well  covered,  it  is  customary  to  slope  the  crown  sheet  of  a  long  furnace 
downwards  towards  the  rear  end,  as  is  shown  in  Fig.  633. 

For  short  furnaces,  such  as  are  generally  found  in  soft-coal  burning  engines,  it  is 
not  deemed  necessaiy  to  slope  the  crown  sheet  unless  the  engine  has  to  run  over  very 
steep  grades. 

But  in  hard-coal  burners,  in  which  the  furnaces  are  always  comparatively  long, 
this  precaution  should  be  taken.  It  may  be  asked,  What  becomes  of  the  front  end  of 
the  crown  sheet  when  the  engine  is  running  up-hill  ?  To  this  we  only  need  to  reply, 
that  the  front  end  of  the  crown  sheet  is  closer  to  the  center  of  the  boiler,  and  conse- 
quently it  cannot  project  above  the  water  as  readily  as  the  back  end  when  the  engine 
is  running  down-hill. 

The  amount  of  slope  should  be  such  as  to  bring  the  crown  sheet  parallel  to  the 
water  level  when  the  engine  is  running  down  the  steepest  grade. 

422.  When  crown  bars  are  used  for  the  purpose  of  staying  the  crown  sheet,  the 
latter  is  generally  made  flat  across  the  furnace,  as  shown  in  Fig.  624.  Sometimes  it  is 
slightly  curved  crossways,  concave  to  the  fire,  having  about  1  \  inches  rise  in  the  center. 
The  advantage  of  this  is  that  the  matter  deposited  from  the  water  tends  to  flow  off,  and 
therefore  prevents  incrustation.  For  some  kinds  of  water  the  crown  bars  cannot  be 
used,  as  they  tend  to  obstruct  the  movements  of  the  matter  deposited  from  the  water, 
and  therefore  tend  to  promote  incrustation,  which  of  course  is  hurtful  to  the  crown 
sheet.  In  such  cases  radial  stay  bolts  are  used,  and  in  order  to  obtain  as  many  perfect 
threads  as  possible  in  the  crown  sheet,  the  latter  is  curved  crossways  to  a  considerable 


LOCOMOTIVE   CONSTRUCTION. 


431 


432 


MODERN  LOCOMOTIVE   CONSTRUCTION, 


extent,  as  is  shown  in  Fig.  634.  This  form  of  crown  sheet  possesses  the  further  advan- 
tage of  being  better  adapted  for  receiving  the  radiant  heat.  But  we  prefer  the  crown 
bars  whenever  the  quality  of  water  will  permit  their  use,  as  we  believe  that  with  crown, 
bars  a  safer  and  stronger  boiler  can  be  made. 


BELPAIKE  FIKE-BOX. 

423.  The  desire  to  avoid  radial  stay  bolts  has  led  to  the  adoption,  on  some  roads, 
of  the  Belpaire  fire-box,  shown  in  Figs.  647  and  648. 

The  difference  between  an  ordinary  and  a  Belpaire  fire-box  is  that  in  the  latter 
the  furnace  crown  sheet  and  the  outer  crown  sheet  are  made  flat  and  placed  parallel  to 

each  other.    The  advantage  of  the  Belpaire  fire-box  is  that  it 
gives  more  steam  space,  and  sometimes  it  is  asserted  that  it 
increases  the  strength  of  the  boiler ;  but  in  this  assertion  we 
/^^^•ffi^fM-ffif$>ii\\     have  no  faith,  for  the  reason  that  the  cylindrical  form  is  the 

n"A^q*rS^£     av" J  |    most  natural  one  to  which  all  vessels  subjected  to  an  internal 

steam  pressure  tend  to  con- 

_     1     j.    L        U   i    i    i    ,.    _ 


form,  and  therefore  the  latter 
must  be  the  strongest. 


ogogo°o°o°o 
o°o°o°ogo°o 


i 

[«       <41  j-outslde-to-outeide-  - 

Pig.  648 


INCLINATION   OF  THE    FURNACE- 
DOOR    SHEETS    AND     SIDE 
SHEETS. 

424.  In  Fig.  655  it  will  be 
seen  that  the  water  space  at 
the  top  of  the  furnace-door 
sheet  is  greater  than  at  the 
bottom — the  difference  is  gen- 
erally about  l£  inches;  the 
aim  is  to  give  the  steam  a 
greater  freedom  of  parting 


Fig.  651 

from  the  sheet.  Experiments  are  recorded  in  which  a  rectangular  metallic  box,  sub- 
merged in  water,  and  heated  from  within,  generated  steam  only  one-half  as  fast  from  its 
vertical  sides  as  it  did  from  the  top,  and  the  bottom  yielded  none.  Again,  by  slightly 
inclining  the  box  the  rate  of  evaporation  on  the  elevated  side  was  increased,  the  steam 
parting  from  it  much  more  easily;  the  steam  on  the  depressed  side  hung  so  slug- 
gishly as  to  lead  to  overheating  of  the  metal.  For  similar  reasons  the  width  of 
the  water  space  towards  the  top  of  the  sides  of  the  furnace  is  also  increased ;  in  fact, 
in  some  boilers  we  find  the  furnace  side  sheets  to  have  a  bulge  only  near  the  flue 
sheet,  and  none  at  rear  end,  where  they  are  straight  and  placed  parallel  to  each  other, 
and  sometimes  even  inclined  towards  each  other  at  the  top.  Although  this  form  of 
fire-box  is  by  no  means  a  universal  one,  we  believe  it  to  be  a  good  one. 


MODERN  LOCOMOTIVE  CONSTRUCTION. 
I 


433 


434 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


WIDTH    OF    WATER    SPACE. 

425.  In  engines  having  cylinders  16  inches  in  diameter  and  upwards,  the  water 
space  at  the  bottom  of  the  rear  end  of  the  fire-box  is  usually  3  or  3£  inches  wide ; 
at  the  bottom  of  the  sides  it  is  also  3  to  3£  inches ;  and  at  the  bottom  of  the  front 
end  it  is  usually  from  3£  to  4  inches  wide. 

For  smaller  engines  the  water  space  is  reduced ;  for  an  engine  with  cylinders  9 
inches  in  diameter  the  water  space  at  the  rear  end  and  sides  is  generally  2£  inches 

wide ;   and  at  the  bottom  of 
the  front  end,  3  inches  wide. 

We  have  seen  the  water 
space  at  the  bottom  of  the  rear 
end  and  the  sides  2  inches 
wide,  but  such  a  narrow  water 
space  should  be  avoided. 


TUBE   SHEETS. 


426.  The  furnace-tube 
sheet  is  sometimes  made  in 
two  pieces,  as  shown  in  Fig. 
633.  With  this  arrangement 
there  is  no  particular  advan- 
tage to  be  gained,  excepting 
the  additional  heating  surface 
in  the  fire-box,  which  is  more 
effective  than  that  of  the 
tubes.  On  the  other  hand, 
there  is  an  objection  to  mak- 
ing the  furnace-tube  sheet  in 

this  way,  as  it  has  been  found  that  the  bend  of  the  flange  at  1)  is  liable  to  wear  and 

give  trouble. 

The  upper  part  of  this  tube  sheet  is  generally  \  inch  thick,  and  the  lower  part 

|  inch  thick.     In  many  engines  the  furnace-tube  sheet  is  made  in  one  piece,  as  shown 

in  Fig.  655 ;  these  sheets  are  generally  J  inch  thick  throughout. 


THICKNESS   OP   FURNACE   SHEETS. 


427.  The  thickness  of  the  furnace-door  and  side  sheets  is  generally  j-6-  inch,  that 
of  the  crown  sheet  is  generally  f  inch ;  and  these  thicknesses  are  used  for  either  iron 
or  steel. 


FIRE-BOX   RING. 


428.  The  ring  or  frame  at  the  bottom  of  the  fire-box  is  made  of  bar-iron,  and  its 
thickness  varies  in  different  engines  from  2  to  2J  inches,  according  to  the  judgment 
and  experience  of  the  different  builders ;  its  width  must,  of  course,  conform  to  the 
water  space. 


MODERN  LOCOMOTIVE  COXSTRrCTIOX. 


435 


The  rivets  are  usually  f  inch  in  diameter,  pass  through  the  rings,  furnace  and  fire- 
box sheets,  and  are  driven  on  the  outside.  At  the  corners  where  there  is  no  room  for 
passing  rivets  through  all  this  metal,  copper  studs  f  inch  in  diameter  are  screwed  into 
the  ring  and  riveted  over  on  the  outer  fire-box  sheets. 


RIVETS   IN   FURNACE   SHEETS. 


429.  In  the  larger  classes  of  engines  all  the  rivets  in   the  furnace   sheets  are 
usually  3  of  an  inch  in  diameter,  and  are  driven  on  the  inside.     It  is  good  practice  to 


•» .IX 


Pig.  060 


Eigf661 


Pigf663 


countersink  the  rivet  holes  in  the  furnace  sheets,  making  the  depth  of  the  countersink 
equal  to  i  of  the  thickness  of  the  sheets,  and  l£  inches  diameter  for  $  rivets.  Sections 
of  these  holes  for  the  various  thicknesses  of  sheets  are  shown  in  Figs.  660,  661,  and 
662.  The  rivets  are  driven  with  points  slightly  raised  above  the  countersink.  Deep 
rivet  heads  in  the  furnace  are  liable  to  burn  off. 


FUSIBLE  PLUGS. 


430.  Fusible  plugs  are  sometimes  used  as  a  safeguard  against  the  collapse  of  the 
furnace  crown  sheet  from  overheating  through  a  shortness  of  water.     (Sections  of  the 


Fig.!  COS 


different  forms  of  these  plugs  are  shown  in  Figs.  663,  664,  and  665.     They  consist  of  a 
brass  shell  containing  an  alloy  of  tin,  lead,  and  bismuth. 

The  plug  is  screwed  into  the  crown  sheet  at  a  distance  of  18  to  24  inches  from  the 
tube  sheet.  The  water  above  the  crown  shoot  keeps  the  alloy  at  a  comparatively  low 
temperature,  and  prevents  it  from  being  melted.  When  the  water  in  the  boiler  is  so 
low  as  to  uncover  the  plug,  the  alloy  is  supposed  to  fuse,  allow  the  stoam  to  escape, 
retard  combustion,  and  in  the  meantime  relieve  the  boiler  of  its  pressure.  Whether 
this  action  of  the  plug  can  always  be  relied  upon  is  very  questionable;  we  believe  its 
efficiency  is  sometimes  over-rated.  A  long  exposure  to  the  heat  in  the  furnace  may 
cause  an  alteration  in  the  nature  of  the  alloy  and  render  it  valueless.  Again,  incrusta- 
tion on  the  plug  may  become  strong  enough  to  withstand  the  pressure  of  the  steam 


4.T) 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


and  prevent  its  escape  after  the  alloy  has  been  melted.  It  is  therefore  good  practice, 
when  these  plugs  must  be  used,  to  renew  them  at  short  intervals,  say  every  two  or 
three  months. 

FIBE-DOOB   OPENINGS. 

431.  In  the  majority  of  locomotives  the  fire-door  opening  is  formed  by  flanging 
the  furnace-door  sheet  outwards,  and  the  back  head  inwards,  bringing  the  seam  in 

the  water  space,  as  shown  in  Fig. 
666.  Sometimes  the  opening  is 
formed  by  riveting  a  flange  to  the 
back  head,  and  a  strip  connecting 
ff'jji"  this  flange  to  the  flange  of  the  fur- 

nace sheet,  as  shown  in  Fig.  667; 
but  this  way  of  forming  the  open- 
ing is  not  as  good  as  the  one  pre- 
viously mentioned.  Occasionally  we 
find  the  opening  made  as  shown  in 
Fig.  668 ;  both  sheets  are  flanged  out- 
wards and  riveted  on  the  outside  of 
the  back  head,  the  same  rivets  taking 
hold  of  a  bar  about  If  or  2  inches 
thick.  This  bar 


*fr^ 


i 


Fig.  668 


extends  around  the  opening ;  it  forms  a  rest  for  the  fire  tools  and  prevents  wear  on 
the  edge  of  the  flange.  It  will  be  noticed  that  the  opening  is  greater  in  the  furnace 
than  it  is  in  tho  boiler  head;  the  object  of  this  is  to  give  the  fireman  a  greater  range 
with  the  fire  tools ;  we  believe  this  to  be  a  good  door  opening.  In  many  locomotives 


MODERJf  LOCOMOTIVE   CONSTRUCTION.  437 

the  fire-door  opening  is  oval  in  form ;  a  good  size  is  14  x  17  inches  in  the  clear ;  of 
course  some  builders  adopt  a  different  size,  but  do  not  vary  much  from  the  size  here 
given.  Sometimes  the  opening  is  made  circular  in  form,  about  16£  inches  in  diameter, 
as  shown  in  Figs.  GG9,  670,  671,  which  also  show  the  construction  of  the  cast-iron 
door  with  its  frame  bolted  to  the  boiler  head.  The  bottom  of  the  opening  should  be 
about  G  inches  above  the  top  of  the  foot-plate. 

DENTS  IN   SIDE  OF  FIRE-BOX. 

432.  Figs.  647  to  653  represent  a  boiler  for  an  engine  which  has  two  axles  below 
the  fire-box.     The  dents  «  and  b  in  the  side  of  the  fire-box  are  for  the  purpose  of 
clearing  the  spring  saddles. 

TUBES. 

433.  The  tubes  are  generally  made  of  iron,  sometimes  of  steel,  seldom  of  copper. 
When  made  of  iron  they  are  manufactured  from  strips  and  lap  welded ;  when  made 
of  steel  they  are  generally  "  solid  drawn."     Steel  tubes  will  probably  be  the  favorite 
ones  in  the  near  future.     They  are  thinner  than  iron  tubes,  and  therefore  the  ends  of 
the  former,  through  a  distance  of  about  6  inches,  are  made  thicker  than  the  body  of 
the  tubes ;  these  ends  are  usually  called  "  safe  ends." 

In  the  ordinary  locomotive  boiler  the  tubes  are  2  inches  external  diameter;  for 
small  boilers  they  are  sometimes  li  inches  external  diameter,  but  it  is  doubtful  whether 
anything  is  gained  by  making  them  so  small.  The  writer  was  once  instructed  to 
design  a  boiler  with  jf-inch  copper  tubes ;  the  boiler  was  made,  but  their  heating 
surface,  on  account  of  the  small  diameter,  was  not  effective ;  they  were  replaced  by 
2-inch  iron  tubes,  and  then  the  boiler  steamed  well. 

When  the  length  of  the  tubes  exceeds  11  feet,  it  is  advisable  to  increase  the 
diameter,  say  to  2J  or  2£  inches.  (See  Art.  437.) 

Formerly  the  thickness  of  iron  flues  2  inches  external  diameter  was  No.  13 
Birmingham  wire-gauge,  but  of  late  the  thickness  has  been  increased  to  about  No.  12 
gauge,  making  the  inside  diameter  equal  to  1.78  inch.  The  thickness  of  2j-inch  tubes 
is  also  about  No.  12  Birmingham  wire-gauge,  making  the  inside  diameter  equal  to 
2.26  inches.  The  thickness  of  IJ-inch  tubes  is  No.  13  wire-gauge,  and  the  inside 
diameter  is  1.31  inch. 

Although  the  primary  object  of  the  tubes  is  to  carry  off  the  gases  and  heat  the 
water  surrounding  them,  they  also  serve  as  stays  for  the  tube  sheets.  Consequently 
many  mechanics  not  only  expand  the  tubes  to  make  them  steam-tight,  but  also  bead 
them  over  to  give  them  a  better  hold  on  the  tube  sheets  so  as  to  prevent  the  steam 
pressure  from  drawing  them  out  of  the  holes. 

Now,  beading  the  tubes  simply  for  the  purpose  of  giving  them  a  better  hold  is 
unnecessary,  because  experiments  have  shown  that  when  a  2-inch  tube  is  properly 
expanded  it  will  resist  a  pull  of  5,000  pounds,  which  is  greater  than  the  stress  1«> 
which  a  locomotive  tube  will  ever  be  subjected.  Hence,  we  frequently  find  the  till ><>s 
not  beaded  on  the  front  flue  sheet;  they  are  simply  expanded  and  allowed  to  project 
J  of  an  inch,  but  they  ,-nv,  or  rather  should  be,  always  beaded  on  the  furnace  tulie 
sheet,  for  the  following  reason : 


438  MODERN  LOCOMOTIVE   CONSTRUCTION. 

The  ends  of  the  tubes  in  the  furnace  are  liable  to  be  exposed  to  cold  currents  of 
air  when  the  furnace  door  is  opened,  which  in  time  will  cause  them  to  leak  and  give 
considerable  trouble.  To  avoid  this  leakage  as  much  as  possible,  copper  ferrules  are 
inserted  between  the  tubes  and  the  sides  of  the  holes  in  the  tube  sheet ;  hence  the  tube 
should  be  beaded  over  to  prevent  the  ferrule  from  dropping  out,  and  at  the  same  time 
protect  the  ferrule  as  well  as  the  end  of  tube  from  wear  caused  by  the  impact  of  the 
uncousumed  fuel  which  is  drawn  against  them. 

Copper  ferrules  are  seldom  placed  on  the  front  ends  of  the  tubes,  although  the 
use  of  ferrules  on  these  ends  we  believe  to  be  good  practice  and  advantageous  when 
the  tubes  have  to  be  taken  out  for  making  repairs,  because  with  ferrules,  the  holes 
through  the  front  tube  sheet,  through  which  the  tubes  must  be  drawn,  are  necessarily 
larger  and  will  permit  a  slight  incrustation  on  the  tubes  to  be  drawn 
through  the  holes,  and  therefore  involve  less  labor  and  time  than  when 
the  holes  ai-e  smaller. 

Without  ferrules,  the  holes  are  drilled  -£%  inch,  and  with  ferrules 
T^  inch  larger  in  diameter  than  that  of  the  tubes. 

434.  Fig.  672  shows  the  ends  of  a  tube  with  a  copper  ferrule.    The 
outside  of  the  ferrule  is  turned  straight,  and  the  inside  is  bored  out 
tapered.     The  length  of  the  feiTule  is  about  J  inch  greater  than  the 
Fig.  era.        thickness  of  the  sheet ;  and  its  greatest  thickness  is  about  &  inch  at 

one  end,  the  taper  bringing  the  opposite  end  to  a  knife-edge. 
The  ordinary  way  of  forming  the  taper  on  the  end  of  the  tube  is  to  hold  it  in  an 
upright  position,  and  allow  it  to  drop  a  few  times  into  a  die  bored  out  to  the  proper 
taper.     Sometimes  the  ferrules  are  bored  out  cylindrical. 

CROSS-SECTIONAL   AREA   OF   TUBES. 

435.  A  too  great  aggregate  cross-sectional  area  of  tubes  produces  a  slow  velocity 
of  the  gases,  which  will  permit  a  deposit  of  soot,  and  cause  a  reduction  in  the  rate  of 
evaporation.  On  the  other  hand,  too  small  an  area  checks  the  draft,  and  besides  this, 
it  leaves  no  allowance  for  a  further  obstruction  to  the  draft  caused  by  the  lodgment  of 
cinders  in  the  lower  tubes,  which  is  liable  to  occur  in  locomotives.  Here,  then,  the 
question  arises :  What  should  be  this  aggregate  cross-sectional  area,"  so  as  to  obtain 
the  best  results  f 

In  Mr.  B.  F.  Isherwood's  "  Experimental  Eesearches  of  Steam  Engineering "  we 
find  that  the  best  aggregate  cross-sectional  area  of  the  tubes  is  /f  of  the  grate  surface ; 
and  it  has  this  property,  that  it  is  the  best  for  all  types  of  boilers,  for  all  rates  of 
combustion,  and  for  all  ratios  of  heating  to  grate  surface. 

The  aggregate  area  may  be  reduced  without  loss  or  much  inconvenience  to  |  of 
the  grate  surface,  but  it  cannot  be  increased  beyond  |  without  serious  sacrifice  to  the 
economic  results.  The  horizontal  fire-tube  boiler  is  much  more  sensibly  affected  by 
this  area  than  the  vertical  water-tube  boiler,  and  requires  a  much  nicer  adjustment  of 
it.  It  is  also  more  affected  by  the  difference  in  ratio  of  the  heating  to  the  grate 
surface,  using  the  same  rate  of  combustion. 

The  aggregate  cross-sectional  area  of  the  tubes  in  American  locomotives  of  the 
larger  sizes  is  about  £  of  the  area  of  the  grate  surface,  and  agrees  with  the  conclusions 


MODEB\  LOCOMOTIVE   CONSTRUCTION. 


439 


of  Mr.  Isherwood ;  but  in  the  smaller  engines  we  sometimes  find  this  area  to  be  as 
la rge  as  i  of  the  grate  surface.  The  reason  for  this  will  be  presently  seen.  The 
following  table  gives  the  aggregate  area  of  tubes  as  found  in  locomotives  running  on 
some  of  our  best  roads. 

TABLE  63. 

TABLE  OP  AGGREGATE  TUBE   AREA   IN   SOFT-COAL  BURNING  ENGINES. 


D'mmetrr  of  Cylin- 
der. 

Stroke. 

Number  of  Tubee. 

Diameter  of 
Tubee. 

Grate  Area. 

Aggregate  Tube 
Area. 

Ratio  of  Orate  to 
Tube  Area. 

Column  1. 

Columns. 

Column  3. 

Column  4. 

Column  5. 

Column  6. 

Column  7. 

Inches. 

Inchw. 

Inchcc. 

Square  feet. 

Square  feet. 

9 

16 

100 

11 

7.08 

1.38 

5.1  to  1 

9 

16 

45 

It 

4.02 

0.62 

6.4  to  1 

10 

18 

110 

H 

8.02 

1.52 

5.2  to 

10 

18 

62 

2 

7.11 

1.06 

6.7  to 

11 

18 

120 

If 

8.50 

1.66 

5.1  to 

12 

20 

130 

2 

13.45 

2.23 

6.0  to 

12 

20 

90 

2 

7.29 

1.55 

4.7  to 

13 

20 

140 

2 

14.16 

2.41 

5.8  to 

14 

20 

145 

2 

14.87 

2.49 

5.9  to 

14 

20 

US 

2 

10.70 

1.96 

5.4  to 

15 

22 

161 

2 

15.58 

2.60 

5.9  to 

15 

22 

126 

2 

11.86 

8.17 

5.4  to 

16 

24 

163 

2 

16.29 

2.80 

5.8  to 

16 

24 

180 

2 

15.28 

3.10 

4.9  to 

17 

24 

184 

2 

17.00 

3.16 

5.3  to  1 

17 

24 

196 

2 

16.28 

3.37 

4.6  to  1 

18 

24 

194 

2 

18.41 

3.34 

5.5  to  1 

18 

24 

200 

2 

17.24 

3.44 

5.0  to  1 

19 

24 

246 

2 

28.00 

4.23 

6.6  to  1 

19 

24 

210 

2 

18.14 

3.61 

5.0  to  1 

20 

24 

252 

2 

31.16 

4.33 

7.1  to  1 

In  examining  the  foregoing  table  our  attention  is  drawn  to  the  fact  that  the 
grate  areas  are  not  proportionate  to  the  size  of  engines ;  indeed,  the  grate  areas  vary 
considerably  for  the  same  size  of  engine.  This  variation  is  partly  accounted  for  by 
the  fact  that  the  proportions  of  boilers  are  not  always  determined  by  fixed  rules,  and 
much  depends  on  the  individual  experience  and  fancy  of  the  designer. 

We  have  already  seen  that  the  aggregate  tube  area  should  be  proportionate  to 
the  grate  area ;  we  shall  therefore  assume  that  the  grate  areas  given  in  Tables  59  and 
GO  are  correct,  and  that  the  ratios  of  these  to  the  tube  areas  are  governed  by  the  follow- 
ing practical  considerations. 

If  the  diameter  and  length  of  a  locomotive  boiler  had  not  been  limited  by  the 
service  for  which  the  engines  is  intended,  then  probably  the  best  proportion  for  the 
total  cross-sectional  area  of  the  tubes  would  be  \  of  the  grate  surface,  which  will 
provide  for  the  loss  of  area  by  the  lodgment  of  cinders  in  the  lower  tubes,  and  leave 
about  ^  of  unimpaired  area.  But  often  we  shall  have  to  be  satisfied  with  an  aggregate 
area  less  than  \  of  the  grate  surface.  For  instance,  in  large  locomotives  the  diameter 
of  the  boiler  is  limited  by  the  distance  between  the  wheels,  which  in  turn  is  limited 
by  the  gauge  of  the  track;  and  if  we  make  the  grate  surfaces  equal  to  the  given  areas 
in  Tables  59  and  GO,  then  it  will  be  difficult  to  obtain  the  desired  number  of  tubes 
without  encroaching  on  the  steam  space,  and  we  will  be  compelled  to  use  a  smaller 
number,  with  a  total  cross-sectional  area  equal  to  about  J  of  the  grate  surface.  If,  ou 


440 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


the  other  hand,  we  make  the  total  area  of  tubes  iu  small  engines  equal  to  \  of  the 
grate  surface,  then  we  shall  have  an  insufficient  heating  surface ;  because  the  wheel 
base  in  these  small  engines  demands  comparatively  short  tubes ;  therefore,  in  order 
to  obtain  a  sufficient  heating  surface  we  must  increase  the  number  of  tubes,  and 
consequently  increase  their  aggregate  cross-sectional  area,  which  in  small  engines  is 
sometimes  equal  to  £  of  the  grate  surface. 

Here,  then,  we  see  the  reason  for  the  variation  of  the  ratio  of  tube  to  grate 
area  in  the  different  classes  of  engines  as  given  iu  Table  64. 

Indeed,  it  may  be  said  that  in  American  locomotives  the  aggregate  cross-sectional 
area  of  the  tubes  varies  from  J  to  J  of  the  grate  surface. 

In  the  following  table  we  have  given  the  number  of  tubes  as  found  by  calculation, 
and  these  will  serve  as  a  guide  in  designing  new  locomotives. 

The  number  of  flues  here  given  agree  closely  with  the  number  used  in  locomotives 
in  actual  service,  and  can  therefore  be  recommended. 

In  Column  3  we  have  given  the  grate  area  as  calculated  by  Rule  103,  reduced  to 
square  inches. 

Column  4  gives  the  ratio  of  the  grate  surface  to  the  aggregate  tube  area;  it 
commences  with  a  ratio  of  4  and  gradually  increases  to  8,  for  the  reason  stated  above. 

Column  5  gives  the  aggregate  cross-sectional  area  of  the  tubes  in  square  inches, 
and  is,  of  course,  obtained  by  dividing  the  grate  area  in  Column  3  by  the  ratio  given 
in  Column  4. 

Column  6  gives  the  number  of  tubes  2  inches  outside  diameter.  The  inner 
diameter  of  these  tubes  has  been  taken  at  1.78  inches,  and  the  cross-sectional  area  of 
one  tube  at  2.48  square  inches ;  therefore  the  number  of  flues  in  Column  6  are  obtained 
by  dividing  the  numbers  in  Column  5  by  2.48. 

TABLE   64. 

TABLE   OP   CALCULATED  AGGREGATE  TUBE  AREA  AND  NUMBER  OF  TUBES. 


Diiimetcr  of  Cylinder. 

Stroke. 

Grate  Area. 

Batio  of  Grate  Area  to 
Aggregate  Tube  Area. 

Aggregate  Cross-sec- 
tional Area  of  Tubes. 

Number  of  Tubes. 

Column  1. 

Column  9. 

Column  8. 

Column  4. 

Column  5. 

Column  6. 

Inches. 

Inches. 

Square  inches. 

Square  inches. 

10 

20 

950.40 

4     to  1 

237.6 

95 

11 

22 

1267.20 

4     to  1 

316.8 

127 

12 

22 

1425.60 

4.5  to  1 

316.8 

127 

13 

22 

1627.20 

5     to  1 

325.4 

131 

14 

24 

1843.20 

5     to  1 

368.6 

148 

15 

24 

2116.80 

5.5  to  1 

384.8 

155 

16 

24 

2275.20 

5.5  to  1 

413.6 

166 

17 

24 

2491.20 

6     tol 

415.2 

167 

18 

24 

2753.20 

6.5  to  1 

423.5 

170 

20 

24 

4320.00 

7.5  to  1 

576.0 

232 

22 

24 

5011.20 

8     to  1 

626.4 

252 

All  tubes  are  2  inches  in  diameter. 


436.  It  may  appear  that  the  foregoing  table  is  only  suitable  for  soft-coal  burning 
engines,  because  in  Column  3  we  have  given  the  grate  surface  for  this  class  of  engines ; 
and  that  for  hard-coal  burning  engines  there  should  be  a  greater  number  of  tubes, 


MODERN  LOCOMOTITE   CONSTRUCTION.  441 

because  in  these  engines  the  grate  surface  is  larger  than  in  soft-coal  burners.  This  to 
some  extent  is  true  ;  but  when  we  remember  that  for  the  smaller  engines  the  total 
area  of  tubes  given  in  the  table  is  greater  than  it  should  be  for  the  given  grate  area, 
and  that  in  the  larger  engines  we  have  not  the  room  for  a  greater  number  of  tubes 
than  given,  we  conclude  that  the  same  number  of  tubes  should  be  used  in  hard-  and 
soft-coal  burners. 

EATIO   OF  DIAMETER  TO  LENGTH  OF  TUBE. 

437.  The  tubes  for  all  the  engines  given  in  Table  64  are  2  inches  external 
diameter,  and  indeed  some  builders  seldom  use  larger  tubes;  but  we  believe  it  to  be 
good  practice  not  to  allow  the  external  diameter  to  be  less  than  -fa  part  of  the  length 
of  the  tubes.     Thus,  for  instance:  if  the  tube  is   12  feet  long,  then   its  external 

12  x  1° 
diameter  should  be  -          -  =  2.2,  say  2J  inches.     Increasing  the  diameter  will  stiffen 

Ot) 

the  tube  and  consequently  prevent  sagging  to  some  extent;  a  comparatively  largo 
diameter  will  also  allow  the  flame  to  extend  further  into  the  tube,  and  make  this 
heating  surface  more  effective.  But  it  must  be  remembered  that  by  increasing  the 
diameter  and  keeping  the  aggregate  cross-sectional  area  the  same  as  for  smaller  tubes, 
the  heating  surface  will  be  decreased.  This  can  be  best  explained  by  taking  an 
example. 

438.  In  Table  64  we  see  that  for  an  engine  with  cylinders  18  inches  diameter 
we  need  170  tubes  2  inches  external  diameter.    The  circumference  of  a  2-inch  tube  is 
•_'  x  3.14  =  6.28  inches,  and  for  170  tubes  the  circumference  is  equal  to  6.28  x  170  = 
1067.60  inches,  or  88.96  feet;   hence  for  every  foot  of  length  of  the  tubes  we  have 
88.96  square  feet  of  heating  surface;  and  the  aggregate  area  of  the  tubes  is  (according 
to  Table  64)  423.1  square  inches.    Now  suppose  we  increase  the  external  diameter  of 
these  tubes  to  2£  inches,  but  leaving  the  length  as  before. 

The  inner  cross-sectional  area  of  one  of  these  tubes  is  4.011  square  inches  ; 
consequently,  since  the  aggregate  area  of  the  tubes  is  not  to  be  changed,  we  have 

4'2l>  1 
-TT  =  105  tubes  2£  inches  diameter  instead  of  170  tubes  2  inches  diameter. 


The  outer  circumference  of  a  2j-inch  tube  is  7.85  inches,  and  the  sum  of  the 
circumference  of  all  the  tubes  is  7.85  x  105  =  824.25  inches,  or  68.68  feet;  hence  for 
every  foot  of  length  of  the  tubes  we  have  68.58  square  feet  of  heating  surface.  But 
the  2-inch  tubes,  with  the  same  aggregate  tube  area,  have  88.96  square  feet,  hence 
by  increasing  the  tubes  to  2i  inches  we  lose  88.96  —  68.68  =  20.28  square  feet  of 
heating  surface  for  every  foot  of  length  of  the  tubes,  which  amounts  to  nearly  23  per 
cent,  loss  of  the  total  tube  heating  surface. 

Although  in  large  boilers  we  may  sacrifice  this  amount  of  heating  surface  for 
other  advantage!  to  be  gained,  we  cannot  afford  to  lose  this  amount  in  small  boilers, 
liccause  in  these  boilers  we  have  had  to  make  the  aggregate  area  of  tubes  larger  than 
desirable  for  the  very  purpose  of  gaining  heating  surface  (see  Art.  435),  and  therefore 
it  would  be  decidedly  bad  practice  to  sacrifice  some  of  this  heating  surface  by  making 
the  diameter  of  the  tubes  greater  than  is  absolutely  necessary.  Therefore  in  small 
boilers  the  diameter  of  the  tubes  should  not  be  greater  than  2  inches,  neither  should 


442 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


it  be  less  than  2  inches,  because  such  a  reduction  will  tend  to  reduce  the  efficiency 
of  the  tube  heating  surface  ;  yet  in  large  boilers  the  diameter  of  the  tubes  may  often 
be  increased,  with  advantage,  to  2£  or  2£  inches. 


LENGTH  OF  TUBES. 

439.  The  length  of  the  tubes  is  limited  by  the  positions  of  the  tube  sheets.     The 
position  of  the  furnace  tube  sheet  is  of  course  established  as  soon  as  the  size  of 
furnace  has  been  determined. 

The  front  tube  sheet  is  placed  as  far  forward  as  possible,  allowing  only  the  heads 
of  the  rivets  through  this  tube  sheet  and  shell  to  clear  the  cylinder  saddle  ;  and  the 
distance  between  these  tube  sheets  will,  of  course,  establish  the  length  of  the  tubes  ; 
since  the  position  of  the  cylinder  saddle  depends  upon  the  length  of  wheel  base,  it 
may  be  said  that  the  wheel  base  determines  the  length  of  tubes,  leaving  us  little  or  no 
choice  in  the  matter. 

ARRANGEMENT  OF  TUBES. 

440.  The  tubes  should  be  arranged  to  allow  the  steam  as  it  is  formed  to  ascend  in 
straight  lines  as  nearly  as  is  possible  for  it  to  do,  therefore  the  favorite  arrangement 
is  like  that  shown  in  Fig.  672a. 

The  tubes  are  usually  spaced  off  on  a  flue  board,  and  from  that  transferred  to  the 
tube  sheets.  The  best  way  to  proceed  in  drawing  the  tubes  on  the  flue  board  will  be 


Fig.  672a. 


Pig.  673 


Pig.  674 


to  establish  first  the  position  of  one  tube  in  the  upper  row,  say  the  tube  a,  Fig.  672«, 
and  through  its  center  draw  the  line  a  e  making  an  angle  of  30  degrees  with  the 
horizontal  line  m  n ;  also,  through  the  center  of  the  tube  a  draw  the  line  a  p  making 
the  same  angle  with  m  n.  On  the  lines  a  e  and  a  p  lay  off  the  centers  of  the  tubes, 
and  through  these  centers  draw  lines  parallel  to  a  e  and  a  p ;  the  intersection  of  these 
lines  will  give  the  centers  of  other  tubes ;  the  centers  above  the  lines  a  e  and  a  p  are 
found  in  a  similar  way.  All  these  lines  can  be  readily  drawn  to  the  correct  angle 
with  the  aid  of  a  T  and  set  square,  the  latter  containing  the  angles  of  30,  60,  and  90 
degrees. 

If  the  tubes  have  been  correctly  located,  then  a  circle  described  from  the  center  of 
any  tube,  say  c,  and  with  a  radius  equal  to  the  distance  between  the  centers  of  tubes, 
will  pass  through  the  centers  of  the  six  nearest  tubes,  asfg  h  ij  k,  shown  in  Fig.  672a, 
When  the  tubes  are  arranged  as  here  shown,  they  are  said  to  be  in  vertical  rows. 
Tubes  may  also  be  arranged  in  vertical  rows  as  shown  in  Fig.  674,  but  this  arrange- 
ment takes  up  too  much  room,  and  therefore  it  is  very  seldom  adopted  in  locomotive 
boilers. 


MODERN  LOCOMOTITK   CONSTRUCTION.  443 

Tubes  placed  as  shown  in  Fig.  673  are  said  to  be  arranged  in  horizontal  rows,  but 
this  arrangement  obstructs  the  upward  passage  of  the  steam,  and  therefore  is  seldom 
adopted. 

441.  The  space  between  the  tubes  depends  sometimes  on  the  number  of  tubes  to 
be  placed  in  the  boiler ;  but  the  best  practice  is  to  make  the  diameter  of  the  boiler  to 
suit  the  number  of  tubes  given  in  Table  64,  and  with  such  spaces  between  them  as 
have  been  found  to  give  good  results.    When  tubes  are  widely  spaced  they  are  more 
easily  cleaned  from  incrustation,  and  there  is  also  less  liability  to  prime.     The  clear 
space  between  two  tubes  should  never  be  less  than  J  inch. 

When  there  are  more  than  120  tubes  in  a  boiler  the  clear  space  between  the  tubes 
should  be  increased  -fV  of  an  inch  for  every  fifteen  additional  tubes.  From  the  fore- 
going we  may  establish  the  following  rule : 

RULE  105. — To  find  the  clear  space  between  the  tubes,  divide  the  number  of  tubes 
in  the  boiler  by  15,  and  multiply  the  quotient  by  -fa  of  an  inch ;  the  product  will  be  the 
required  clear  space  between  the  tubes.  Add  to  this  the  diameter  of  the  tube ;  the 
sum  will  be  the  distance  between  the  centers.  This  rule  will  hold  good  only  when 
thore  are  more  than  120  tubes  in  the  boiler.  For  a  smaller  number  of  tubes  the  space 
should  always  be  \  inch. 

EXAMPLE  140. — What  should  be  the  distance  from  center  to  center  of  tubes  in  a 

135 
boiler  having  135  tubes  2  inches  diameter  ?    Here  we  have  -rp-  =  9 ;  and  9  x  -^  =  fa 

inch  for  the  space  between  the  tubes.  Adding  this  to  the  diameter  of  the  tube,  we 
have:  &  +  2  =  2  fa  inches  from  center  to  center  of  tubes. 

In  a  similar  way  we  find  that  the  space  between  the  tubes  in  a  boiler  with  195 

195 
tubes  should  be  -yp  =  13 ;  and  13  x  fa  =  ^f  inch.    If  the  tubes  are  2i  inches  outside 

diameter,  then  the  distance  from  center  to  center  should  be  2£  -I-  f£  =  3^  inches. 
We  believe  that  the  circulation  will  be  improved  by  arranging  the  two  central  vertical 
rows  of  tubes,  as  shown  in  Fig.  626,  and  making  the  clear  space  between  them  from 
J  to  £  inch  greater  than  the  space  between  the  other  tubes.  Although  this  arrange- 
ment is  sometimes-  adopted,  it  is  not  a  general  practice ;  they  are  usually  arranged 
as  shown  in  Fig.  656. 

STEAM   SPACE. 

442.  The  distance  from  top  of  the  upper  row  of  flues  to  the  top  of  the  boiler 
varies  from  J  to  J  of  the  diameter  of  the  boiler ;  we  prefer  J,  because  a  liberal  steam 
space  tends  to  prevent  priming ;  it  must  also  be  remembered  that  when  the  tubes  are 
placed  too  high  the  water  surface  is  contracted,  which  by  itself  is  probably  a  greater 
cause  for  priming  than  simply  a  contracted  steam  space.     In  all  cases  the  steam  space 
should  be  large  enough  to  allow  a  man  to  enter  the  boiler  to  make  repairs  when 
necessaiy. 

DIAMETER  OF  BOILER   SHELL  AND  DESIGN  OF  BOILER. 

443.  Having  established  the  number  and  size  of  tubes.  :ils<>  th<>  height  of  steam 
space,  we  are  in  a  position  to  determine  the  diameter  of  boiler,  which  has  simply  to 


444  MODERN  LOCOMOTIVE   CONSTRUCTION. 

be  made  large  enough  to  admit  the  required  number  of  tubes  with  sufficient  steam 
space  above  them,  and  with  a  clearance  of  not  less  than  2  inches  between  any  tube 
and  the  boiler  shell. 

We  have  now  sufficient  data  for  determining  the  principal  dimensions  of  a 
locomotive  boiler  ;  and  from  the  foregoing  remarks  we  conclude  that  the  proper  mode 
of  procedure  in  designing  a  boiler  is  as  follows  : 

First.  Find  the  grate  area  as  explained  in  Art.  414. 

Second.  Find  the  aggi-egate  cross-sectional  area  of  the  tubes  according  to  Art.  435. 

Third.  Divide  the  aggregate  area  of  the  tubes  by  the  inner  cross-sectional  area 
of  one  tube  ;  the  quotient  will  be  the  required  number  of  tubes. 

Fourth.  Locate  the  tubes  —  that  is  to  say,  make  an  end  view  of  the  tubes  placed 
at  the  proper  distance  apart,  and  draw  an  end  view  of  the  boiler-shell  which  will 
admit  the  required  number  of  tubes,  leaving  a  steam  space  equal  to  £  of  the  diameter 
of  the  shell  in  height,  and  a  2-inch  space  between  the  side  and  bottom  tubes  and 
shell. 

Fifth.  Make  an  outline  drawing  of  the  boiler,  an  end  view,  and  a  longitudinal 
section,  and  see  to  it  that  there  is  ample  clearance  between  the  boiler  and  the 
mechanism  attached  to  it,  and  make  provision  for  fastening  the  boiler  to  the  frames. 

EXAMPLE  141.  —  Find  the  principal  dimensions  of  a  boiler  for  an  eight-wheeled 
passenger  engine  having  cylinders  16  inches  diameter  and  24  inches  stroke  ;  driving 
wheels,  58  inches  diameter  ;  mean  effective  steam  pressure,  90  pounds  per  square  inch  ; 
soft-coal  burner  ;  tubes,  2  inches  outside  diameter. 

First.  According  to  Eule  103,  we  divide  the  tractive  force  by  600  ;  the  quotient 
will  be  the  grate  surface  in  square  feet.  The  tractive  force  of  this  engine  is 
(according  to  Rule  3)  9,533  pounds  ;  hence  we  have 

9533 

=  15.8  square  feet. 


Second.  According  to  Table  64,  the  aggregate  area  of  the  tubes  for  this  size  of 

1  15.8  x  144 

engine  should  be  ^7  part  of  the  grate  area  ;  we  therefore  have  -  '   ,  -          =  414.4 

O.O  O.O 

square  inches  for  the  aggregate  area  of  the  tubes. 

Third.  To  find  the  number  of  tubes.  Referring  to  Art.  435,  we  see  that  the  inner 
cross-sectional  area  of  a  tube  2  inches  outside  diameter  is  2.48  square  inches  ;  hence, 
dividing  the  aggregate  area  of  the  tubes  by  the  cross-sectional  area  of  one  tube,  we  have 

413.6 
2  AQ   =  166,  which  is  the  required  number  of  tubes.     Referring  to  Art.  441,  we  see 

1  C*  C* 

that  the  clear  space  between  the  tubes  should  be  -^  —  11,  and  11  x  ^  =  H  of  an 


inch,  and  2  +  y^  =  2y^  inches  from  center  to  center  of  tubes. 

Fourth.  To  find  the  diameter  of  boiler  shell.  Lay  in  about  200  tubes  2f£  inches 
from  center  to  center,  and  then  find  by  trial  a  circle  which  will  take  in  166  tubes, 
leaving  a  clear  space  of  about  £  of  its  diameter  in  height  above  the  upper  row  of  tubes, 
and  not  less  than  2  inches  clearance  between  the  side  and  bottom  tubes  and  shell  ; 
the  diameter  of  this  circle  will  be  the  inside  diameter  of  shell.  It  may  be  remarked 


MODERN   LOCOMOTIVE    CONSTRUCTION.  445 

here  that  the  grate  surface,  cross-sectional  tube  area,  and  number  of  tubes  in  this 
example  could  have  been  found  at  once  without  any  calculation  by  referring  to  Table  64 

Fifth.  A  longitudinal  section  and  end  view  of  the  boiler  should  now  be  drawn, 
and  all  such  mechanism  as  may  come  in  contact  with  boiler  should  be  located  and 
examined  in  regard  to  clearance,  and  allowances  made  for  fastening  the  boiler  to  the 
frames. 

If  the  fire-box  has  to  be  placed  between  the  driving  axles,  it  may  be  that  on 
account  of  insufficient  room  the  grate  area  of  15.8  square  feet,  as  previously  found, 
will  be  somewhat  too  large,  and  will  have  to  be  reduced.  We  shall  need  about  6 
inches  from  the  center  of  the  rear  driving  axle  to  rear  end  of  fire-box,  and  about  13  to 
14  inches  from  the  center  of  the  front  axle  to  the  front  end  of  the  fire-box ;  but  the 
reduction  of  the  grate,  if  any,  will  be  comparatively  small,  so  that  no  change  will  need 
to  be  made  in  the  number  of  tubes. 

HEATING  SURFACE. 

444.  The  heating  surface    throughout   the  boiler  is  not  equally  effective;  the 
relative  value  will  depend  on  its  position.     A  square  foot  of  heating  surface  in  the 
furnace  will  evaporate  considerably  more  water  than  a  square  foot  of  tube  surface ; 
and  a  square  foot  of  heating  surface  in  the  rear  end  of  the  tubes  will  be  more  effective 
than  a  square  foot  in  the  front  end ;  the  tops  of  the  flues  are  more  effect- 
ive than  the  bottoms.     The  relative  value  of  the  different  parts  of  a  tube 

as  a  heating  surface  is  generally  represented  as  follows : 

Let  Fig.  675  represent  the  end  of  a  tube.  If  now  the  efficiency  of 
the  upper  part  of  the  tube  is  considered  to  be  equal  to  1,  the  efficiency  of 
the  sides  will  be  equal  to  4,  and  the  value  of  the  bottoms  will  be  0.  Hence  the  average 

efficiency  of  the  whole  tube  surface  will  be  -  -  =  £.    This  shows  that  the 

whole  surface  of  the  tube  is  only  half  as  effective  as  it  would  be  if  all  the  different 
parts  had  been  as  valuable  as  the  top. 

445.  Experiments  also  indicate  that  the  heating  surface  placed  perpendicular  to 
the  current  of  heated  gases,  so  as  to  receive  the  heat  by  direct  impact,  is  more  effective 
than  a  heating  surface  placed  diagonally  to  the  current;  approximately,  one  square 
foot  of  the  former  is  as  effective  as  4  square  feet  of  the  latter,  and  it  will  require  8 
square  feet  of  heating  surface  placed  parallel  to  the  current  to  transmit  as  much  heat 
as  1  square  foot  placed  perpendicular  to  the  current. 

446.  In  computing  the  heating  surface  of  a  locomotive  boiler  the  difference  of  the 
efficiency  is  not  taken  into  account,  and  the  whole  surface  is  treated  as  if  there  existed 
no  difference. 

All  the  surface  above  the  top  of  grate  bars,  in  contact  with  the  water  and  trans- 
mitting heat  to  it,  is  considered  to  be  heating  surface. 

Many  builders  endeavor  to  obtain  400  square  feet  of  heating  surfuro  for  every 
cubic  foot  of  piston  displacement  during  one  stroke.  If,  for  instance,  the  cylinders  are 
16  inches  diameter  and  24  indies  stroke,  the  piston  displacement  in  one  cylinder  will 
be  2.79  cubic  feet,  and  consequently  the  total  heating  surface  should  bo  2.79  x  400  = 


446  MODERN  LOCOMOTIVE   CONSTRUCTION. 

1,116  square  feet.  But  in  many  eases  the  wheel  base  and  gauge  of  track  will  limit  the 
heating  surface,  and  we  may  have  to  be  satisfied  with  320  square  feet  per  cubic  foot  of 
piston  displacement. 

Some  master-mechanics  take  into  account  the  diameter  of  the  wheel  and  mean 
effective  pressure  per  square  inch  of  piston,  and  consequently  proportion  the  heating 
surface  to  the  tractive  force,  allowing  from  1  square  foot  of  heating  surface  for  every  10, 
up  to  12  pounds  of  tractive  force.  Thus,  for  instance,  the  tractive  force  of  an  eight- 
wheeled  passenger  engine  with  cylinders  16  x  24  inches,  mean  effective  steam  press- 
ure 90  pounds  per  square  inch,  and  driving  wheels  58  inches  diameter,  will  be  9,533 
pounds.  Now,  allowing  1  square  foot  of  heating  surface  for  every  10  pounds  of  tractive 

c\r"  oo 

force,  we  have  -^  =  953  square  feet  of  heating  surface,  which  is  less  than  found 
before. 

If  we  allow  1  square  foot  of  heating  surface  for  every  10  pounds  of  tractive  force, 
and  1  square  foot  of  grate  surface  for  every  600  pounds  of  tractive  force,  then  the 
ratio  of  heating  to  grate  surface  will  be  as  60  to  1,  which  agrees  well  with  the  average 
practice.  The  ratio  of  heating  to  grate  surface  varies  from  45  to  1  to  70  to  1. 

For  hard-coal  burning  engines  1  square  foot  heating  is  also  allowed  for  eveiy  10 
pounds  of  tractive  force ;  and  allowing  1  square  foot  of  grate  surface  for  every  500 
pounds  of  tractive  force,  the  ratio  of  heating  to  grate  surface  will  be  50  to  1,  and  this 
again  agrees  well  with  the  average  practice. 

The  heating  surface  in  the  tubes  is  from  6  to  8  times  greater  than  the  heating 
surface  in  the  fire-box ;  probably  a  fair  average  will  be  8  to  1,  so  that  when  the  total 
heating  is  953  square  feet,  that  in  the  tubes  will  be  847.2  square  feet. 

447.  The  foregoing  proportion  will  aid  us  in  estimating  the  length  of  the  tubes. 
For  example,  let  it  be  required  to  find  the  length  of  the  tubes  for  an  eight- wheeled 
passenger  engine,  with  16  x  24  inch  cylinders ;  diameter  of  driving  wheels,  58  inches ; 
mean  effective  steam  pressure,  90  pounds  per  square  inch. 

The  tractive  force  of  this  engine  is  9,533  pounds ;  its  total  heating  surface  will  be 

-^Q-  =  953  square  feet ;  and  the  heating  surface  in  the  tubes  will  be  953  -  -^   =  847.2 

square  feet.  According  to  Table  64,  we  find  that  this  engine  should  have  166  tubes  2 
inches  outside  diameter.  The  circumference  of  a  2-inch  tube  is  6.28  inches,  and  for 
166  tubes  the  sum  of  the  circumferences  will  be  166  x  6.28  =  1042.48  inches,  or  86.87 
feet.  Dividing  the  required  tube  heating  surface  of  847.2  square  feet  by  86.87  feet,  we 

847.2 
have  ggg_  =  9.7  feet  for  the  length  of  the  tubes.     Whether  we  can  use  tubes  of  this 

length  will  depend  on  the  wheel  base. 

The  value  of  the  tube  heating  surface  is  often  overrated,  but  still  it  has  value, 
and  designers  should  strive  to  obtain  such  a  length  for  the  tubes  as  to  give  a  liberal 
heating  surface. 

KIVETED  JOINTS. 

448.  The  common  practice  in  American  locomotive  boiler  construction  is  to  use 
single-riveted  lap  joints  for  the  circular  seams,  and  double-riveted  lap  joints  for  the 
longitudinal  seams ;  or  a  double-riveted  lap  joint  for  the  circular  seams,  and  double- 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


447 


riveted  lap  joints  with  welt  pieces  for  the  longitudinal  seams.  Sometimes  welt  pieces 
are  also  put  over  the  circular  seams  up  to  the  water  line,  for  the  purpose  of  preventing 
furrowing.  Sometimes  the  longitudinal  seams  are  butt  joints. 

The  single-riveted  lap  joint  is  represented  in  Fig.  676.     It  is  the  simplest  kind  of 
riveted  joints.     A  double-riveted  lap  joint  is  represented  in  Fig.  677.     The  latter  joint 


Fiff. 


Jt 


I) 


Fig.  677 


is,  of  course,  stronger  than  the  former.  The  reason  for  using  a  stronger  joint  for  the 
longitudinal  seams  is,  that  every  inch  of  this  seam  has  to  resist  twice  the  stress  of 
every  inch  in  the  circular  seain.  To  prove  this  we  will  take 
the  following  illustration : 

Let  Fig.  678  represent  a  cylinder  36  inches  diameter,  one 
inch  long,  and  conceive  the  shell  to  be  very  thin,  so  that  for 
the  purpose  of  investigation  we  may  leave  the  thickness  out 
of  consideration ;  also  let  the  steam  pressure  in  the  cylinder 
be  100  pounds  per  square  inch. 

The  stress  tending  to  part  the  sheet  on  any  longitudinal  line, 
as  a  i,  will  be  equal  to  one-half  the  product  obtained  by  multiplying  the  diameter,  d  a, 
by  the  steam  pressure  per  square  inch ;  hence  we  have 


36  x  100 


=  1800  pounds, 


which  is  the  stress  on  the  line  a  b  or  d  c,  and  since  these  lines  are  one  inch  long,  we  may 
say  that  this  is  the  stress  per  inch  of  the  longitudinal  seam.  The  total  stress  on  any 
circular  line  will  be  equal  to  the  area  of  the  head  in  square  inches  multiplied  by  the 
steam  pressure  per  square  inch.  The  area  of  the  head  is  1017.88  square  inches,  hence 
the  total  pressure  on  the  head  will  be  1017.88  x  100  =  101,788  pounds.  But  this 
pressure  is  resisted  by  the  circumference  of  the  cylinder,  which  is  equal  to 

36  x  3.1416  ==  113.09+  inches. 

Here,  then,  we  have  113.09  inches  to  resist  101,788  pounds,  consequently  each  inch 
will  have  to  resist 

101788 

11309  =        0  pounds, 

which  is  just  one-half  of  the  stress  per  inch  on  the  longitudinal  lino  a  b. 

This  can  be  proved  in  another  way.     Lot  p  denote  the  steam  pressure  per  square 
inch ;  d,  the  diameter  in  inches  of  a  shell  one  inch  long ;   7j,  the  tension  or  stress  per 


448 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


inch  on  the  longitudinal  seam ;  Tc,  the  tension  per  inch  on  the  circular  seam.     Then 
for  the  longitudinal  seam  we  have 

T  ==  d  x  P 

O         ? 

and  for  the  circular  seam  we  have 

d"-  x  .7854 
=  d  x  3.1416  X  p- 

This  hist  equation  can  be  written  as  follows : 


By  canceling  we  have 


=  d  x  d  x  .7854 
•  ~  d  x  4  x  .7854  > 


But 


T  = 

-*-  a 


_  d  x  p 


4 
d  x  p 


which  shows  the  tension  or  stress  per  inch  of  the  circular  seam  is  just  one-half  of  that 
on  one  inch  in  the  longitudinal  seam.  But  from  the  foregoing  it  must  not  be  under- 
stood that  the  double-riveted  lap  joints,  when  used  in  a  boiler  for  the  longitudinal 
seams,  are  twice  as  strong  as  the  single-riveted  lap  joints  used  for  the  circular  seams. 


EFFECTS   OF   TENSION   ON   EIVETED   JOINTS. 

449.  When  a  riveted  joint,  is  subjected  to  tension,  fracture  may  occur  in  several 
ways,  depending  on  the  defects  of  construction. 

Let  the  shaded  portion  in  Fig.  679  represent  a  strip  of  a  boiler  shell,  with  a  single- 
riveted  lap  joint,  and  let  the  width  a  be  equal  to  the  pitch  of  the  rivets.  It  is  evident 
that  the  strip  is  held  together  by  the  section  of  the  metal  whose  width  is  equal  to  b ; 
that  is,  the  cross-sectional  area  of  the  metal  between  the  rivets ;  it  is  also  held  together 
by  the  cross-sectional  area  of  one  rivet.  Now,  fracture  may  occur  in  the  following  ways : 

First.  By  the  metal  breaking  across  in  front  of  the  rivet,  as  shown  at  c,  Fig.  679, 
the  action  being  similar  to  that  of  the  fracture  of  a  beam  loaded  at  the  center  and  fixed 


.A. 

Fig.  679 


Fig.  680  Fiff'  681 

at  the  ends.  The  cause  of  failure  lies  in  the  fact  of  punching  or  drilling  the  rivet 
holes  too  close  to  the  edge  of  the  plate.  It  is  therefore  of  great  importance  to  place 
the  rivets  at  a  proper  distance  from  the  edge  of  the  plate. 


MOVERS   LOCOMOTIVE   CONSTRUCTION. 


449 


Prof.  W.  C.  Unwin,  in   his  "  Elements  of  Machine   Design,"  gives  the  following 
distances  from  the  center  of  rivets  to  the  edges  of  iron  and  steel  plates : 

TABLE  65. 


Diameter  of  Rivets. 

Iron  Plates. 

Steel  Plates. 

Distance  from  Edge  of  Plate 
to  Center  Line  of  Rivets. 

Distance  from  Edge  of  Plate 
to  Center  Line  oi  Biveu. 

4  in 

f 

eh. 

.00  in 
.14 

sh. 

0.86  in 
0.98 

ch. 

i 

1 

.29 
.41 
.55 

1.12 
1.22 
1.35 

H 
li 

.67 
L.80 

1.46 
1.57 

Mr.  R.  Wilson,  in  his  "  Treatise  on  Steam  Boilers,"  says :  "  The  lap  (that  is,  from 
edge  to  edge  of  plates)  for  single  riveting  should  be  equal  to  3  times  the  diameter  of 
the  rivet,  and  never  more  than  3.3  times  the  diameter." 

The  ordinary  practical  rule  is  to  make  the  distance  from  center  line  of  rivets  to 
the  edge  of  plate  equal  to  lj  times  the  diameter  of  the  rivet.  For  thin  plates,  up  to 
4  inch  thick,  we  would  recommend  to  make  this  distance  If  times  the  diameter  of  the 
rivet  for  iron  and  steel  plates. 

If  the  lap  is  too  great,  there  will  be  difficulty,  owing  to  the  elasticity  of  the  metal, 
in  calking  the  joint  so  as  to  make  it  steam-tight. 

Second.  Fracture  may  occur  by  crushing  the  plate  and  rivet,  as  shown  in  Fig.  680, 
causing  the  joint  to  become  leaky  and  insecure.  But  if  the  diameter  of  the  rivet  is  of 
the  proportion  to  thickness  of  plate  as  given  in  Table  66,  trouble  of  this  kind  need  not 
be  anticipated. 

Third.  Fracture  may  occur  by  shearing  the  rivet,  as  shown  in  Fig.  681. 

This  is  caused  by  placing  the  rivets  too  far  apart,  or  using  rivets  too  small  in 
diameter.  Here,  then,  the  following  question  presents  itself :  What  shall  be  the  diame- 
ter of  the  rivets  ?  This  question  leads  us  to  consider  the  diameter  of  the  smallest  hole 
which  can  be  punched  through  a  plate  of  given  thickness.  In  punching  the  hole,  the 
area  sheared  by  the  punch  is  equal  to  -ndt — in  which  it  denotes  the  ratio  of  the  diam- 
eter to  the  circumference,  and  is  always  equal  to  3.1416 ;  d,  the  diameter  of  hole ;  and 
t,  the  thickness  of  plate,  all  in  inches.  Now  let  c,  denote  the  resistance  to  shearing  per 
square  inch  of  metal,  and  let  It,  be  the  total  resistance  to  shearing ;  then  we  have 

R,   =  TrdtCf. 

The  strength  of  the  punch  is  equal  to  its  cross-sectional  area  multiplied  by  the  resist- 
ance to  crushing  per  square  inch.  Let  ce  represent  the  resistance  to  crushing  per 
square  inch ;  and  Ee  the  whole  resistance  to  crushing.  Remembering  that  the  cross- 
sectional  area  of  the  puin-h  is  equal  to  jrf2,  we  have 


450 


MODEKN  LOCOMOTIVE   CONSTRUCTION. 


And  when  the  total  resistance  to  crushing  is  equal  to  the  total  resistance  to  shearing, 
we  have 


=  irdtc 


, 


=  4:ndtct. 


or, 

clearing  of  fractions,  we  have 
Hence, 


Therefore  the  diameter  of  the  punch  can  never  be  less  than  — * ;  if  it  is  less,  the 

punch  will  be  crushed.  If,  for  instance,  the  thickness  t  of  the  plate  is  \  inch,  and  the 
resistance  c,  to  shearing  is  20  tons  per  square  inch,  and  the  resistance  cc  to  crushing  is 
80  tons  per  square  inch,  then  by  substituting  for  the  letters  their  values  in  the 

formula,  we  have 

4  x  4  x  20 
d  =  -   —^ —    ~  =  $  inch 

for  the  smallest  diameter  of  the  punch  which  can  be  used  for  a  plate  £  inch  thick ;  or, 
as  will  be  seen  in  this  particular  case,  the  diameter  of  the  punch  is  equal  to  the  thick- 
ness of  the  plate. 

In  practice,  of  course,  the  diameter  of  the  rivet  holes  is  always  larger  than  the 
thickness  of  the  plate,  so  as  to  reduce  the  pressure  per  square  inch  on  the  punch. 

In  practice  the  diameter  of  the  rivet  is  generally  equal  to  J  of  the  thickness  of  the 
plate  plus  §  of  an  inch.  Or,  if  d  denotes  the  diameter  of  the  rivet  in  inches,  and  t  the 
thickness  of  the  plate  in  inches,  we  may  give  in  place  of  the  foregoing  rule  the  sym- 
bolic expression, 

d  =  %t  +  |  of  an  inch. 

The  diameters  of  the  rivets  in  Table  66  have  been  found  by  the  foregoing  rule. 
The  first  column  gives  the  thickness  of  the  plates,  and  the  second,  the  diameters  of  the 
rivets  in  the  nearest  ^  of  an  inch  found  by  calculations. 

TABLE  66. 


Thickness  of  Plate. 

Diameter  of  Rivets. 

ft  inch. 

^   inch. 

i 

ft 

A 

H 

A 

* 

i 

It 

ft 

1 

1 

H 

H 

i 

s 

lA 

TS 

it" 

Fourth.  Fracture  may  occur  along  the  center  line  x  x  of  rivets,  Fig.  679.  It  is 
caused  by  placing  the  rivets  too  close  to  each  other.  This  leads  us  to  the  consideration 
of  the  pitch  of  rivets — that  is,  the  distaijce  between  their  centers. 


MODERN   LOCOMOTITE    CONSTRUCTION.  451 

The  tensile  strength  of  iron  plates  in  the  direction  of  the  grain  varies  from  about 
4J,0(H)  to  50,000  pounds  per  square  inch.  Now,  for  the  purpose  of  calculating  the 
pitch  of  rivets,  we  should  know  the  exact  tensile  strength ;  sometimes  it  is  stamped  on 
the  plates,  but  when  this  is  omitted  we  should  cut  testing  strips,  and  find  the  tensile 
strength  by  actual  tests.  In  the  absence  of  this  knowledge  we  shall  assume  that  the 
tensile  strength  is  45,000  pounds  per  square  inch. 

Referring  to  Fig.  679,  let  the  shaded  portion  represent  a  strip  of  boiler  shell ;  the 
width  a  of  this  strip  is  equal  to  the  pitch  of  the  rivets. 

Along  the  center  line  x  x  of  the  rivets  the  width  of  the  strip  is  necessarily  reduced 
to  the  width  6,  and  the  amount  of  this  reduction  is  equal  to  the  diameter  of  the  rivet. 
Now,  if  a  square  inch  of  metal  can  resist  45,000  pounds  at  a,  it  would  appear  that  a 
square  inch  of  metal  will  also  resist  45,000  pounds  at  b.  But  this  is  not  true ;  the 
reasons  for  this  statement  are  as  follows:  First,  referring  to  /?,  Fig.  679,  wo  notice  that 
the  line  d  e,  which  represents  the  direction  in  which  the  pull  or  tension  acts,  does  not 
pass  through  the  center  of  the  thickness  of  those  portions  of  the  plates  which  make  up 
the  riveted  joint ;  hence  the  action  of  the  tension  on  the  joint  is  oblique,  and  there- 
fore the  plates  at  the  joint  cannot  resist  as  great  a  pull  as  they  can  directly  above  or 
below  the  joint. 

Second.  Punching  will  reduce  the  tenacity  of  the  plate  along  the  line  of  rivet 
holes.  The  probable  cause  of  the  reduction  of  tenacity  or  injury  is,  that  in  punching 
the  metal  immediately  surrounding  the  hole  is  squeezed  laterally  into  the  plate,  giving 
this  portion  of  the  metal  a  permanent  set,  thereby  reducing  the  strength  of  the  plate. 
We  believe  that  with  a  spiral  punch  the  injury  will  not  be  as  great  as  when  the  plates 
are  punched  with  a  common  flat-ended  punch.  The  recorded  results  of  experiments 
made  for  determining  the  amount  of  reduction  of  strength  caused  by  punching  vary 
so  much  that  it  is  impossible  to  state  the  correct  amount.  But  experiments  seem  to 
indicate  that,  if  the  holes  are  punched  somewhat  too  small,  so  that  a  ring  of  about  .06 
of  an  inch  or  more  can  be  reamed  out  to  bring  the  holes  to  the  right  size,  the  injury 
to  the  plate  will  practically  amount  to  nothing ;  or,  if  the  holes  are  punched  to  the 
right  size,  and  the  plates  are  annealed  after  punching,  the  original  tenacity  will  be 
restored 

In  the  following  calculations  we  shall  assume  that  the  plates  are  not  annealed,  and 
that  the  holes  are  punched  to  the  right  size  with  an  ordinary  flat-ended  punch,  all  of 
which  is  a  common  practice.  And  we  shall  further  assume  that  the  oblique  action  of 
the  pull  on  the  joint,  combined  with  the  loss  due  to  punching,  reduces  the  original 
strength  of  45,000  pounds  per  square  inch  to  40,000  pounds  per  square  inch;  the 
latter  amount  we  shall  hereafter,  for  the  sake  of  brevity,  call  the  apparent  strength. 

It  is  reasonable  to  assume  that  in  new  boilers  properly  made  friction  exists 
between  the  parts  of  the  plates  which  make  up  the  joints,  and  this  friction  adds 
strength  to  the  joints.  But  the  contraction  and  expansion  will  in  time  reduce  this 
friction  to  nothing;  we  shall  therefore  neglect  it  in  computing  the  strength  of  the 
joint.  Now,  if  the  apparent  strength  of  the  plate  at  the  joint  is  40,000  pounds  per 
square  inch,  as  we  have  assumed,  it  will  be  evident  that,  if  we  multiply  40,000  pounds 
by  the  area  of  the  section  at  b — that  is,  multiply  40,000  pounds  by  the  thickness  of 
the  plate,  and  by  the  width  6,  we  obtain  the  total  stress  which  our  shaded  strip  in  Fig. 


452  MODERN  LOCOMOTIVE   CONSTRUCTION. 

679  can  resist,  provided  the  rivet  is  strong  enough.  Let  Rr  represent  the  total  resist- 
ance of  the  plate  at  b ;  and  let  the  distance  in  inches  between  the  sides  of  the  rivets  be 
denoted  by  b;  t,  the  thickness  in  inches  of  the  plate;  and  cn  the  apparent  strength 
per  square  inch,  which  is  equal  to  40,000  pounds.  We  then  have 

Rr  =  bxtxcr.  (1) 

Since  the  width  a  of  the  shaded  strip  in  Fig.  679  is  equal  to  the  pitch  of  the 
rivets,  it  follows  that  only  one  whole  rivet  can  hold  the  two  parts  of  the  strip  together. 
Now  it  must  be  readily  perceived  that,  in  order  to  make  one  part  of  the  joint  as 
strong  as  the  other,  one  rivet  must  be  capable  of  resisting  just  as  great  a  pull  as  the 
plate  at  its  section  b. 

In  a  lap  joint  the  rivets  are  subjected,  to  a  small  extent,  to  an  oblique  shearing 
stress,  and  it  is  perhaps  for  this  reason,  and  on  account  of  the  bending  action  of  the 
plates,  that  the  rivets  do  not  offer  as  great  a  resistance  to  shearing  as  they  would  do 
when  shearing  takes  place  in  a  plane  perpendicular  to  the  axis  of  the  rivets.  Again, 
rivets  in  punched  holes  seem  to  offer  a  greater  resistance  to  shearing  than  those  in 
drilled  holes.  The  difference  is  probably  due  to  the  sharp  edges  of  the  drilled  holes, 
which  seem  to  shear  the  rivets  with  less  pressure  or  load  than  the  blunter  edges  of  a 
punched  hole.  We  shall  assume  that  the  shearing  resistance  of  iron  rivets  in  punched 
holes  is  46,000  pounds  per  square  inch.  The  total  shearing  resistance  of  the  rivet 
will  then  be  equal  to  its  cross-sectional  area  multiplied  by  46,000.  Now,  putting  Re  for 
the  total  shearing  resistance  of  the  rivet,  d,  the  diameter  of  the  rivet  in  inches,  and 
cs  for  the  shearing  resistance  per  square  inch  =  46,000  pounds,  we  have 

Re  =  .7854  x  (P  x  c,.  (2) 

But  since  the  total  shearing  resistance  of  the  rivet  must  be  equal  to  the  strength 
of  the  plate  at  its  section  b  (see  Fig.  679),  we  have  R,  =  Rr;  and  according  to 
Equation  1,  Er  =  b  x  t  x  cr;  in  which  b  is  the  distance  between  the  sides  of  the 
rivets.  Now,  this  distance  b  is  evidently  equal  to  the  pitch  of  the  rivets  minus  the 
diameter  of  the  rivet.  Putting  p  for  the  pitch  and  d  for  the  diameter,  we  have 

Rr  =  (p  —  d)  x  t  x  cr. 
Since  R,  =  Rr,  we  have 

.7854  x  d2  x  c,  =  (p  —  d)  x  t  x  cr. 
From  the  foregoing  we  get 

.7854  x  d2  x  c, 

~T^T    •  +  *-*• 

But  .7854  x  d2  is  equal  to  the  cross-sectional  area  of  the  rivet ;  if  we  now  put  a 
for  this  area,  we  have 

a  x  c, 

-  +  d  —  p 
t  x  cr 

Replacing  c,  and  cr  by  their  values  as  previously  assumed,  we  have 

a  x  46000 


t  x  40000 


d=p.  (105) 


MODERN   LOCOMOTIVE   CONSTRUCTION. 


453 


EXAMPLE  141«. —  Find  the  pitch  of  i-ivets  in  a  single-riveted  lap  joint  for  an  iron 
plate  £  of  an  inch  thick. 

In  Table  66  we  find  that  the  diameter  of  the  rivet  for  a  $-inch  plate  should  be 
Ij^e  inches.  The  cross-sectional  area  of  the  rivet  is  .8866  square  inch.  Now,  substitut- 
ing for  the  symbols  in  Formula  105  their  values,  we  have 


.8866  x  46000 
.75  x  40000' 


-I-  If^  =  p  =  2118  inches. 


In  this  manner  we  have  found  the  pitch  of  rivets  in  the  nearest  fa  of  an  inch 
given  in  Column  3,  Table  67.  It  will  be  noticed  that  this  pitch  does  not  increase 
regularly  with  the  thickness  of  the  plates ;  it  is  therefore  not  adopted  in  practice. 
The  pitch  given  in  Column  4  (same  table)  is  recommended,  as  it  agrees  closer  with 
the  average  practice. 

In  Table  67  it  will  be  noticed  that  the  pitch  of  rivets  as  recommended  for  thin 
plates  is  less  than  the  calculated  pitch.  The  reason  for  this  reduction  is  to  avoid  the 
difficulty  in  calking.  On  the  other  hand,  for  thick  plates  the  pitch  of  rivets  as  recom- 
mended is  greater  than  the  calculated  pitch.  The  reason  for  this  increase  is  twofold : 
First.  In  Formula  105  we  have  taken  the  nominal  diameter  of  the  rivet,  not  the  real 
diameter — that  is  to  say,  the  holes  are  always  punched  somewhat  greater  than  the 
nominal  diameter  of  the  rivet ;  this,  of  course,  decreases  the  amount  of  metal  between 
the  rivets ;  and  since  the  rivet  generally  fills  the  holes,  the  real  diameter  of  the  rivet 
is  larger  than  its  nominal  diameter,  and  for  this  increase  no  allowance  was  made  in  the 
formula,  hence  we  see  the  propriety  of  increasing  the  calculated  pitch.  Second.  The 
edges  of  the  plates  will  corrode  or  wear  faster  than  the  rivets.  This  fact  indicates 
that  the  calculated  pitch  should  be  increased,  provided  it  does  not  interfere  with 
calking,  and  since  thick  plates  are  not  so  liable  to  spring,  a  limited  increase  of  the 
calculated  pitch,  as  given  in  Column  4,  can  be  made  without  endangering  the  tight- 
ness of  the  joint.  Without  this  increase  of  pitch  the  life  of  the  boiler  is  shortened. 
Of  course,  for  thin  plates  an  increase  of  the  calculated  pitch  is  prevented  by  the 
difficulty  of  calking. 

TABLE   67. 

PITCH    OF    KIVF.TS    FOK   SINGLE-RIVETED   LAP  JOINTS   FOB   WROUGHT-IRON   PIATES   PUNCHED 

AM)   NOT   ANNEALED. 


Thickness  of  Plates. 

Diameter  of  Rlveta. 

Pitch  KB  Calculated  by 
Formula  1U5. 

Pilch  as  Recommended. 

Column  1. 

Column  s. 

Column  8. 

Column  4. 

•fg  inch. 

4    inch. 

1  !  ,',  inches. 

\^  inches. 

| 

!  ',. 

1J 

1* 

ft 

I,1 

-i' 

2 

1 

24 

24 

13 

-  r'i 

24 

4 

}J 

24 

2| 

ft 

f« 

si 
i 

;-;< 

U 

1 

21 

2f 

ft 

1   ,', 

2i 

It 

1 

3 

1 

14 

24 

3 

454 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


Many  engineers  assume  the  resistance  of  the  plate  between  the  rivet  holes  to  be 
equal  to  the  shearing  resistance  of  the  rivets.  Under  these  conditions  the  formula  for 
finding  the  pitch  of  rivets  will  be  simpler  than  Formula  105.  It  is  as  follows : 

a 

-  +  d=p,  (106) 

in  which  a  denotes  the  cross-sectional  area  of  the  rivet  in  square  inches ;  t,  the  thick- 
ness of  the  plate  in  inches ;  d,  the  diameter  of  the  rivet  in  inches ;  and  p,  the  pitch  in 
inches. 

The  pitch  found  by  this  formula  will  be  less  than  that  given  in  Table  67.  We 
have  stated  that  the  shearing  resistance  of  the  rivets  in  drilled  holes  is  not  as  great 
as  that  of  the  rivets  in  punched  holes,  consequently  when  the  holes  are  drilled  it  is 
advisable  to  reduce  the  pitch  which  is  given  in  Table  67,  and  adopt  the  pitch  given  in 
the  following  table,  which  has  been  calculated  by  the  formula, 


This  pitch  is  also  more  suitable  for  steel  plates  and  iron  rivets,  or  in  all  cases  where 
the  tensile  strength  of  the  plate  and  the  shearing  strength  of  the  rivets  are  equal  or 
nearly  so. 

TABLE   68. 

a 

PITCH  OP  RIVETS  FOR  SINGLE-RIVETED  LAP  JOINTS,  CALCULATED  BY  FORMULA    ,    +  «  =  p ;    SUITABLE 
FOR   WROUGHT-IRON   PLATES   WITH   DRILLED   HOLES,   OR   STEEL   PLATES   WITH   IRON    RIVETS. 


Thickness  of  Plates. 

Diameter  of  Rivets. 

Pitch  of  Rivets  as  Cal- 
culated by  Formul-i 

Pitch  of  Rivets  Recom- 
mended. 

Column  1. 

Column  2. 

Column  3. 

Column  4. 

-1%  inch. 

^    inch. 

1  |!1,j  inches. 

1^  inches. 

i 

ft 

Ift 

H 

A 

u 

1J 

Hi 

1J 

J^ 

la 

2 

2 

J 

« 

2 

8* 

ft 

i 

m 

24 

t 

It 

2  ,-j; 

2f 

ft 

i 

24 

'    2i 

ift 

2$ 

4 

ift 

2ft 

2f 

H 

* 

2J 

DOUBLE-RIVETED   LAP   JOINTS. 

450.  We  will  now  consider  double-riveted  lap  joints.  Let  the  shaded  portion, 
Fig.  682,  represent  a  strip  whose  width  b  is  equal  to  the  pitch  of  a  double-riveted  lap 
joint.  Here  it  is  evident  that  two  rivets  hold  the  joint  together.  In  all  other  respects 
the  conditions  are  just  the  same  as  those  for  a  single-riveted  lap  joint,  and  the  remarks 
relating  to  it  given  in  Art.  44-9  are  also  applicable  to  the  double-riveted  lap  joint. 


LOCOMOTIVE 


455 


Since  there  are  two  rivets  in  the  shaded  strip  (Fig.  C82),  the  formula 

«  x  46000 


t   x  40000 


d  = 


as  given  for  single-riveted  joint,  must  be  slightly  changed  so  as  to  adapt  it  for  a  double- 
riveted  joint.  It  is  as  follows : 

2a  x  46000 

TT-40000  +  d  =  A  (107) 

in  which  a  denotes  the  cross-sectional  area  in  square  inches  of  the  rivet;  t,  the  thick- 
ness of  the  plate  in  inches ;  rf,  the  diameter  of  the  rivet  in  inches ;  and  p,  the  pitch  in 
inches. 

The  pitch  given  in  Column  3  in  Table  69  has  been  calculated  by  the  foregoing 
formula. 

TABLE   69. 

PITCH   OP   RIVETS    FOR   A   DOUBLE-RIVETED   LAP   JOINT,    WROUGHT-IRON   PLATES,    PUNCHED  AND 

NOT   ANNEALED. 


Thickness  of  Plates. 

Diameter  of  Rivets. 

Pitch  of  Riveta  aa  Cal- 
culated by  Formula. 

Pitch  of  Riveta  Rec- 
ommended. 

Dlniance  C  between 
the  Rows  of  KivetB. 

Column  1. 

Column  •-'. 

Column  8. 

Column  4. 

Column  6. 

,*6  inch. 

i   inch. 

21  inches. 

2  inches. 

H  inches. 

J 

ft 

3|" 

2i 

4 

1 

i 

gf 

8* 

A 

3i 

2t 

H 

i 

TI 

3i 

3 

14 

rfe 

I 

3| 

3i 

it 

* 

iJ 

U 

i* 

H 

1 

5| 

3f 

A 

IA 

IA 

3ft 

3f 
3J 

A 

i 

H 

3| 

4 

2 

The  pitch  recommended  in  Column  4  is,  of  course,  the  distance  P  from  center  to 
center  of  rivets  measured  on  the  center  line  of  one  row  of  rivets ;  the  distance  C  given 
in  Column  5  is  the  perpendicular  distance  between  the  two  rows  of  rivets  (Fig.  682). 


Fig.  6S2 


The  distances  C  have  been  adjusted  to  suit  the  conditions  shown  graphically  in 
Fig.  683.  In  this  figure  we  have  divided  the  metal  between  t  li«-  siilcs  of  any  two  rivets, 
k  and  /,  for  instance,  in  four  equal  parts,  and  from  the  center  of  rivets  A,  /,  and  i 


456 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


described  circles  whose  diameters  are  equal  to  the  diameter  of  one  rivet  plus  two  of 
the  four  equal  parts  laid  off  between  the  rivets.  Now,  lines  drawn  perpendicular  to 
the  line  k  I  through  the  points  of  division  1  and  3  will  be  tangent  to  the  circles  around 
the  centers  k  and  I,  and  drawing  lines  perpendicular  to  k  I  and  tangent  to  the  opposite 
sides  n  and  o  of  these  circles,  and  also  drawing  lines  in  the  same  direction  and  tangent 
to  the  circumference  of  the  rivets,  we  may  assume  that  these  lines  represent  straps  y  r/ 
pulling  on  the  rivets  k  and  I ;  and  we  may  further  assume  that  an  eye  bar  /  is  pulling 
on  the  rivet  i.  In  order  to  make  the  eye  of  the  bar  /  as  strong  as  the  straps  g  y,  the 
outlines  of  the  eye  bar  /  and  the  straps  must  not  overlap  each  other ;  in  fact,  to  be  on 
the  safe  side  a  small  space  should  be  left  between  the  circles  n  1  around  k,o  3  around 
I,  and  h  m  around  i.  Hence  the  distances  C  given  in  all  the  tables  have  been  adjusted 
so  that  the  rivets  in  the  upper  row  will  fall  outside  of  the  arcs  d  e  and  d  <?2  (Fig.  682), 
whose  radii  are  equal  to  half  the  pitch  drawn  from  centers  of  two  adjacent  rivets  in  the 
lower  row. 

If  we  assume  that  the  apparent  tenacity  of  the  plates  between  the  rivets  is  equal 
to  the  shearing  resistance  of  the  rivets,  then  Formula  107  will  become 


—  +  d  =  p, 


(108) 


in  which  a  represents  the  cross-sectional  area  of  the  rivet  in  inches ;  t,  the  thickness 
of  the  plate  in  inches ;  d,  the  diameter  of  the  rivet  in  inches ;  and  p,  the  pitch  in 
inches  as  before. 

From  this  formula  the  pitch  given  in  Column  3,  Table  70,  has  been  calculated. 

The  pitch  recommended  in  this  table  we  believe  to  be  more  suitable  for  iron  plates 
with  drilled  holes,  or  steel  plates  with  iron  rivets,  than  the  pitch  given  in  Table  69. 

TABLE  70. 

PITCH   OF  RIVETS  FOR  DOUBLE-RIVETED  LAP  JOINTS,  CALCULATED  BY  FORMULA   —  +  d  =  p,   SUIT- 
ABLE  FOR  WROUGHT-IRON   PLATES   WITH   DRILLED   HOLES   OR   STEEL   PLATES  WITH   IRON   RIVETS. 


Thickness  of  Plates. 

Diameter  of  Rivets. 

Pitch  of  Rivets  as  Cal- 
culated by  Formula. 

Pilch  of  Rivets  Rec- 
ommended. 

Distance  C  between 
the  Rows  of  Rivets. 

Column  1. 

Column  2. 

Column  3. 

Column  4. 

Column  6. 

ft  inch. 

^   inch. 

2f  inches. 

2  inches. 

H   inches. 

i 

ft 

2ft 

8* 

1ft 

ft 

H 

8A 

2i 

11 

1 

4 

si 

2i 

1ft 

ft 

H 

3ft 

21 

If 

i 

ii 

3i 

2f 

1* 

ft 

1 

3 

21 

It 

i 

« 

3ft 

3 

If 

H 

i 

3i 

84 

Iff 

1A 

3ft 

3i 

If 

n 

ift 

3ft 

3| 

lit 

* 

ii 

3f 

3i 

2 

It  should  not  be  understood  that  the  pitch  for  single  and  double-riveted  lap  joints 
as  given  in  the  foregoing  tables  must  be  used  in  all  cases.  But  when  the  tenacity  of 
the  plates  and  the  shearing  resistance  of  the  rivets  are  not  known,  we  believe  that 
the  pitch  given  in  the  tables  is  as  good  as  can  be  adopted.  On  the  other  hand, 


MODERN  LOCOMOTIVE  CONSTRUCTION.  457 

when  the  tenacity  of  the  plates  and  the  shearing  resistance  of  the  rivets  are  known, 
the  pitch  should  be  determined  by  the  given  formulas,  taking  care  to  substitute  the 
proper  values  for  the  symbols  cr  and  c,,  and  making  such  allowances  for  calking  and 
wear  of  plates  as  in  the  judgment  of  the  designer  are  necessary. 


STRENGTH  OF  RIVETED  JOINTS   COMPARED  WITH  THAT  OF  THE   SOLID  PLATE. 

451.  Let  us  now  compare  the  strength  of  a  single-riveted  lap  joint  with  the 
strength  of  a  solid  plate.  Referring  to  Table  67,  we  find  in  Column  4  that  the  pitch 
recommended  for  a  plate  -fa  inch  thick  is  l£  inches,  and  in  Column  3  we  notice  that 
in  order  to  make  the  plate  between  the  rivets  as  strong  as  the  rivets,  the  pitch 
should  have  been  !-}-&  inches;  hence  we  conclude  that,  by  adopting  l£  inches  for  the 
pitch,  the  plate  will  be  weaker  than  the  rivets,  and  therefore,  for  comparing  the 
strength  of  the  lap  joint  with  the  strength  of  the  solid  plate,  we  only  need  to  take 
into  consideration  the  strength  of  the  plate  between  the  rivets. 

In  a  plate  of  this  thickness  the  diameter  of  the  rivets  is  %  inch.  Now,  assuming 
that  the  rivet  holes  are  punched  Vu  of  an  inch  larger  in  diameter,  we  have  for  the 
width  of  the  metal  between  the  rivets  l£  —  -£$  =  If  =  .937  inch.  Multiplying  this 
width  by  the  thickness  &  =  .187  inch,  we  have  .937  x  .187  =  .175+  square  inch  for 
the  area  of  the  metal.  According  to  Art.  449,  the  apparent  strength  of  the  metal  at  b 
(see  Tig.  679)  is  40,000  pounds  per  square  inch,  hence  the  resistance  to  tearing  of  the 
metal  between  the  rivets  is  .175  x  40000  =  7000  pounds. 

The  width  a  of  the  solid  strip  (Fig.  679)  is  equal  to  the  pitch,  l£  inches ;  the 
cross-sectional  area  of  this  strip  is  1.5  x  .187  =  .2805  square  inch.  We  have  assumed 
that  the  tenacity  of  the  plate  above  and  below  the  joint  is  45,000  pounds,  hence  the 
resistance  to  tearing  of  the  solid  part  of  the  plate  is  .2805  x  45000  =  12622.5  pounds. 
Therefore  the  strength  of  the  single-riveted  lap  joint  compared  with  the  strength  of 
the  solid  plate  is 

7000  x  100 
12622.5      =  55  per  cent< 

Now  let  us  take  a  plate  J  inch  thick ;  for  this  thickness  the  pitch  recommended 
in  Column  4,  Table  67,  is  3  inches,  but  in  Column  3  in  the  same  table  we  find 
that  the  pitch  should  have  been  2£  inches,  hence  we  conclude  that  by  adopting  a 
pitch  of  3  inches,  the  metal  between  the  rivets  is  stronger  than  the  rivets;  con- 
sequently, in  comparing  the  strength  of  the  joint  with  that  of  the  solid  plate,  we 
only  need  to  take  the  strength  of  the  rivet  into  consideration. 

Assuming,  again,  that  the  rivet  holes  have  been  punched  ^  of  an  inch  larger  in 
diameter  than  the  rivets,  the  diameter  of  the  rivet  when  properly  driven  becomes  1  ,a0  in- 
stead of  l£,  as  given  in  the  table.  The  cross-sectional  area  of  a  rivet  !,-„  inches 
diameter  is  1.107  square  inches,  and  for  the  shearing  strength  we  have  adopted  46,000 
pounds  per  square  inch,  therefore  the  total  shearing  resistance  of  the  rivet  is  1.107  x 
46000  =  50922  pounds.  The  width  a  of  the  strip  (Fig.  679)  is  equal  to  tin-  pitch,  3 
inches ;  multiplying  this  by  the  thickness  I  =  .875  inch,  we  have  3  x  .875  =  2.625 
square  inches  for  the  cross-sectional  area  of  the  strip.  The  tenacity  of  the  plate  we 


458 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


have  taken  at  45,000  per  square  inch,  consequently  the  resistance  to  tearing  of  the 
solid  part  of  the  strip  will  be 

2.625  x  45000  =  118125  pounds. 

And  the  strength  of  the  joint  compared  with  that  of  the  solid  plate  will  be 

50922  x  100 


118125 


=  43  per  cent. 


Calculating  by  the  foregoing  methods  the  strength  of  the  metal  between  the  rivets, 
and  also  the  strength  of  the  rivets  for  each  thickness  of  plate,  and  comparing  the 
weaker  of  the  two  with  the  strength  of  the  solid  plate,  we  obtain  the  percentages  given 
in  the  following  table,  and  these  represent  the  efficiency  of  the  joints  in  which  the  pitch 
is  equal  to  that  recommended  in  Table  67. 

TABLE  71. 

EFFICIENCY   OP   SINGLE-RIVETED   LAP  JOINTS,    IRON   PLATES,   PUNCHED   HOLES,   NOT   ANNEALED. 


Thickness  of  Plates. 

DiametiT  of  Rivets. 

Pitch  of  Rivels. 

Efficiency  of  Joints. 

Column  1. 

Column  2. 

Column  3. 

Column  4. 

V\  inch. 

i   inch. 

1J  inches. 

55  per  cent. 

i 

ft 

If 

2 

57 
55 

t 

24 

54 

ft 

If 

2* 

54 

if 

2| 

53 

W 

i 

2^ 

50 

i 

H 

2f 

49 

H 

i 

24 

47 

l 

i^L 

2^ 

47 

Ift 

3 

41 

i 

14 

3 

43 

When  the  tenacity  of  the  plates,  apparent  strength  of  metal  between  the  rivets, 
and  the  shearing  strength  of  rivets  differ  from  the  values  which  we  have  assumed, 
then  of  course  the  efficiency  of  the  joints  will  also  differ  from  that  given  in  the  fore- 
going table,  but  with  the  proper  selection  of  the  material  for  the  boiler  the  difference 
will  not  be  very  great. 

Mr.  Fairbairn  gives  56  per  cent,  for  the  efficiency  of  a  single-riveted  joint.  This 
agrees  very  closely  with  the  efficiency  of  thin  plates  given  in  our  table,  but  it  seems 
that  for  thick  plates  56  per  cent,  is  too  great. 

In  a  way  similar  to  the  foregoing  we  find  the  efficiency  of  a  double-riveted  lap 
joint.  The  results  are  given  in  the  following  table ;  they  have  been  calculated  for  the 
pitch  given  in  Table  69. 


MODERX  LOCOMOTIVE   CONSTRVCTION. 


459 


TABLE   72. 

EFFICIKNCY  OF   DOUBLE-RIVETED   LAP  JOINTS,   IRON   PLATES,   PUNCHED   HOLES,    NOT  ANNEALED. 


Thickness  of  Plates. 

Diameter  of  Rivets. 

Pitch  of  Rivet*. 

Efficiency  of  Joints. 

Column  1. 

Column  'J. 

Column  8. 

I  '011111111  4. 

A  inch. 

i  inch. 

2  inches. 

63  per  cent. 

f. 

ft 

2i 
24 

62 
59 

,1 

i 

W 

2i 
2* 

59 
60 

i 

ft 

3 

61 

ft 

i 

3i 

63 

H 

3i 

63 

H 

i 

3* 

62 

I 

1A 

3f 

62 

H 

IS 

3i 

63 

H 

4 

63 

This  table  shows  us  that  for  the  assumed  values  of  the  tenacity  of  the  plates, 
shearing  strength  of  rivets,  etc.,  the  efficiency  of  double-i'iveted  lap  joints  varies  from 
60  to  63  per  cent.  Mr.  Faii-bairn  gives  70  per  cent,  for  the  efficiency  of  these  joints. 

In  Art.  448,  we  have  intimated  that  the  double-riveted  lap  joint  has  less  than  twice 
the  strength  of  the  single  joint.  We  are  now  in  a  position  to  state  how  much  stronger 
one  is  than  the  other.  Referring  to  Table  72,  we  may  say  that  the  average  strength 
of  the  double-riveted  joint  is  62  per  cent,  of  that  of  the  solid  plate ;  and  referring  to 
Table  71,  we  may  say  that  the  average  strength  of  the  single-riveted  joint  is  49  per 
cent,  of  that  of  the  solid  plate.  Consequently  the  double-riveted  joint  is  1.26,  say  1J 
times  as  strong  as  the  single-riveted  joint ;  and  we  will  find  the  same  difference  by 
adopting  Mr.  Fairbairu's  values. 

RIVETED  JOINTS  WITH  WELT  PIECES. 

452.  Fig.  684  shows  an  excellent  longitudinal  seam ;  it  is  stronger  than  ordinary 
double-riveted  lap  joints,  and  differs  from  the  latter  by  having  a  welt  piece  added  on 
the  inside,  extending  along  the  whole  length  of 
the  joint.  The  welt  piece  makes  the  joint  stiffer, 
and  therefore  will  somewhat  increase  the  ap- 
parent strength  of  metal  between  the  rivets,  con- 
sequently the  pitch  of  the  rivets  in  this  joint  can 
and  should  be  made  a  little  less  than  that  given 
in  the  table ;  by  so  doing,  the  difficulty  of  draw- 
ing the  three  thicknesses  together  will  be  reduced. 

The  inside  welt  piece  not  only  adds  strength 
to  the  joint,  but  it  also  serves  another  very  useful 
purpose,  namely,  it  prevents  furrowing — that  is, 
the  wearing  or  eating  away  of  the  metal  on  the  inside  of  the  boiler  dose  to  the  edges 
of  the  joint,  forming  in  some  cases  clear  and  distinct  grooves.  As  to  the  cause  of  fur- 
rowing, engineers  do  not  fully  agree,  but  the  weight  of  evidence  seems  to  indicate  that 


460  MODERN  LOCOMOTIVE   CONSTRUCTION. 

furrowing  occurs  to  a  greater  extent  with  the  use  of  bad  water  than  with  the  use  of  good 
water.  Again,  furrowing  occurs  only  along  the  seams  below  the  water-line,  and  seldom, 
if  ever,  above  it.  This  seems  to  indicate  that  there  must  be  a  chemical  action  of  the 
water  on  the  iron.  The  cause  is  also  attributed  to  the  form  of  the  lap  joint.  Eeferring 
to  Fig.  679,  we  notice  that  the  line/*/  of  the  lap  is  not  in  line  with  the  center  line  d  e  of 
the  thickness  of  the  shell ;  consequently,  when  the  boiler  is  subjected  to  a  steam  press- 
ure, the  tendency  will  be  to  draw  the  line  /  g  to  coincide  with  the  line  d  e,  causing  a 
disturbance  in  the  fibers  of  the  iron  near  the  edge  of  the  plates,  exposing  a  raw  place 
to  the  chemical  action  of  the  water,  and  thereby  promoting  furrowing ;  hence  it  is  some- 
times said  that  furrowing  is  caused  by  a  combined  chemical  and  mechanical  action. 
But  whatever  may  be  the  cause,  the  fact  remains  that  furrowing  does  occur  below  the 
water-line,  and  not  above  it.  Locomotive  boilers  made  of  -f$  inch  iron  have  been  known 
to  furrow  nearly  through  the  thickness  of  the  iron  in  eighteen  months.  On  the  other 
hand,  boilers  with  welt  pieces  have  been  known  not  to  furrow  during  twenty-five  years 
of  hard  service.  This  indicates  an  absolute  necessity  of  placing  all  the  longitudinal 
seams  above  the  water-line  whenever  it  is  possible  to  do  so,  and  that  all  transverse 
seams  should  have  welt  pieces  extending  to  above  the  water-line.  Indeed,  this  has 
become  the  practice  on  roads  where  the  water  is  bad. 

In  the  report  of  the  proceedings  of  the  American  Railway  Master-Mechanics' 
Association,  an  instance  is  recorded  where  furrowing  was  prevented  by  putting  in  an 
old  boiler  a  sheet  -£•$  inch  thick,  extending  about  one-third  around  the  circle  of  the 
barrel  and  throughout  its  whole  length,  and  riveted  lightly  to  the  barrel. 

This  sheet  or  liner  was  first  placed  in  the  old  boiler  for  the  purpose  of  strengthen- 
ing it ;  but  when  the  boiler  was  examined  after  about  two  or  three  years'  service,  it 
was  found  that  furrowing  had  not  continued. 

From  the  foregoing  facts  we  learn  that  welt  pieces  not  only  increase  the  strength 
^  |   I          of  boilers,  but  they   also  add  to  their  life ;  and  we  may  conclude  that 

all  longitudinal  seams  in  any  boiler   should  have  welt  pieces;  below 
(®)         the  water-line  the  welt  pieces  are  needed  for  strength  and  to  prevent 

furrowing,  above  the  water-line  they  are  needed  for  strength. 

For  similar  reasons  all  single-riveted   transverse  or  circular  seams 

should  have  welt  pieces  all  around ;  and  double-riveted  transverse  seams 

should  have  welt  pieces  extending  to  above  the  water-line. 

453.  Fig.  685  shows  a  butt  joint  having  a  welt  piece  inside  as  well 
as  outside  of  the  shell.  This  joint  is  not  so  often  used  in  this  country,  but  it  is 
frequently  adopted  in  Europe,  where  it  is  the  favorite  one. 


THICKNESS   OF  BOILER   SHELL. 

454.  It  is  scarcely  necessary  to  say  that  the  seam  is  the  weakest  part  of  the  boiler, 
and  the  weakest  part  must  be  made  strong  enough  to  resist  the  pressure. 

In  Tables  71  and  72  we  have  given  the  strength  of  joints  compared  with  that  of 
the  solid  plate.  This  comparison  will  now  enable  us  to  compute  the  thickness  of  the 
shell  for  any  diameter  and  steam  pressure. 

EXAMPLE  142. — It  is  required  to  find  the  thickness  of  the  plate  in  a  steel  boiler  50 


0) 


MODERN  LOCOMOTIVE  CONSTRUCTION.  461 

inches  inside  diameter,  subjected  to  a  steam  pressure  of  150  pounds  per  square  inch. 
The  ultimate  tensile  strength  of  the  steel  is  assumed  to  be  65,000  pounds  per  square  inch. 
Longitudinal  seams  are  to  be  ordinary  lap  joints  double-riveted  with  welt  pieces,  holes 
punched,  plates  not  annealed  after  punching. 

In  example  of  this  kind  we  need  to  take  the  strength  of  the  longitudinal  seams 
only  into  account  ;  if  these  are  strong  enough,  the  circular  seams  will  also  be  strong 
enough,  because  the  stress  per  square  inch  on  the  longitudinal  seams  is  twice  as  great 
as  that  on  the  circular  seams,  but  the  former  are  not  twice  as  strong  as  the  latter. 

For  the  sake  of  simplicity  let  us  employ  for  the  relative  strength  of  the  seam  to 
that  of  the  solid  plate  the  value  given  by  Mr.  Fairbairn,  which  for  a  double-riveted 
joint  is  70  per  cent,  of  the  solid  plate  (Art.  451). 

The  ultimate  strength  of  the  solid  plate  as  given  in  the  example  is  65,000  pounds 
per  square  inch  ;  hence  the  strength  of  the  seam  is  found  by  the  following  proportion  : 

100  :   70  :  :  65000  :  x. 
Working  out  this  proportion,  we  have 

70  x  65000 

..«..  —       =  45500  pounds  per  square  inch 

of  the  plate  in  the  seam.  But  this  45,000  pounds  per  square  inch  is  the  ulti- 
mate strength  of  the  seam  —  that  is  to  say,  if  the  joint  is  subjected  to  this 
stress  it  will  break.  The  general  custom  is  to  subject  the  joint  to  not  more  than  £  of 
its  ultimate  strength,  or,  in  other  words,  a  factor  of  safety  of  5  is  adopted  ;  hence  the 

limit  of  the  stress  to  which  the  joint  should  be  subjected  is  —  ^—  =  9100  pounds  per 

D 

square  inch,  and  this  value  (9100  pounds  per  square  inch)  can  always  be  employed  for 
double-riveted  joints  with  welt  pieces,  in  plates  whose  ultimate  tensile  strength  is 
65,000  pounds  per  square  inch  ;  we  shall  therefore  call  this  value  a  constant.  Now  the 
total  stress  to  which  a  strip  one  inch  wide  of  the  longitudinal  joint  will  be  subjected  is 
numerically  equal  to  the  inner  radius  in  inches  of  the  boiler  multiplied  by  the  steam 
pressure  per  square  inch.  Let  P  denote  the  steam  pressure  in  pounds  per  square  inch, 
and  It  the  inner  radius  in  inches  of  the  boiler  ;  then  the  total  stress  to  which  a  strip 
of  the  seam  one  inch  wide  will  be  subjected  is  equal  to  P  x  R. 

The  stress  which  a  strip  one  inch  wide  of  the  joint  can  resist  with  safety  is  numer- 
ically equal  to  the  thickness  of  the  plate  multiplied  by  the  constant  previously  found. 
Let  T  be  the  thickness  of  the  plate  in  inches  ;  C,  the  constant  ;  then  the  resistance  which 
a  1-inch  strip  of  the  joint  can  offer  with  safety  is  equal  to  T  x  C.  Since  a  strip  one 
inch  wide  has  to  resist  a  stress  equal  to  P  x  It,  we  have 

T  x  fj  =  P  x  It. 
From  this  we  obtain 

T    p*x. 
c 

Substituting  for  the  symbols  their  values,  we  have 


,)1()() 
which  is  the  thickness  of  the  plate. 


.      x  IT) 
1  •         ,)()()        •  .11'-'  inch,  say  ^  of  an  inch, 


462  MODERN  LOCOMOTIVE  CONSTRUCTION. 

In  working  out  this  example,  we  have  assumed  that  the  strength  of  the  joint  is  70 
per  cent,  of  that  of  the  solid  plate ;  this  value  is  somewhat  too  high  for  a  double-joint 
without  welt  pieces — its  strength  is  probably  not  over  60  per  cent,  of  the  solid  plate 
(see  Table  72).  But  if  welt  pieces  are  used,  as  they  should  be  in  all  boilers,  then  the 
value  of  70  per  cent,  for  the  joint  is  not  too  great  and  may  be  safely  adopted. 

In  a  similar  way  we  may  find  the  thickness  of  plates  for  boilers  of  any  diameter, 
but  it  must  be  remembered  that  the  value  of  the  constant  C  changes  with  the  kind  and 
quality  of  the  material  employed,  and  with  the  kind  of  joint  adopted. 

For  the  sake  of  convenience  we  give  the  following  formulas  for  computing  the 
thickness  of  plates  for  the  ordinary  class  of  boiler  work,  and  for  computing  the 
pressure  which  ordinary  boilers  made  of  iron  whose  original  ultimate  tensile  strength  is 
45,000  pounds  per  square  inch,  and  those  made  of  steel  with  a  tenacity  of  65,000  pounds 
per  square  inch  will  stand.  These  tenacities  we  believe  to  be  about  the  average 
in  ordinary  boiler  work.  The  factor  of  safety  is  5,  and  for  the  relative  strength  of 
joints  we  have  adopted  Mr.  Fairbairn's  values. 

For  iron  boilers  with  single-riveted  longitudinal  lap  joints,  welt  pieces  inside : 

P  x  R  T  x  5040 

5040  '  R 

For  steel  boilers  with  single-riveted  longitudinal  lap  joints,  with  welt  pieces  inside : 

P  x  R  T  x  7280 

7280  '  R 

For  iron  boilers  with  double-riveted  longitudinal  lap  joints,  with  welt  pieces  inside : 

P  x  R  T  x  6300 

6300  '  R 

For  steel  boilers  with  double-riveted  longitudinal  lap  joints,  with  welt  pieces  inside : 

P  x  R  T  x  9100 

9100  '  R 

In  all  these  formulas  T  denotes  the  thickness  of  the  plate  in  inches ;  P,  the  steam 
pressure  per  square  inch ;  and  It,  the  inside  radius  in  inches  of  the  barrel. 


STEEL,   IRON,   AND  COPPEK  FIRE-BOXES. 

455.  Iii  Europe  copper  is  used  to  a  great  extent  for  fire-boxes.  In  this  country 
experience  indicates  that  steel  is  decidedly  the  best  material  for  this  purpose,  although 
some  master-mechanics  still  use  iron  and  believe  that  it  gives  as  good  results  as  steel. 
The  failui'e  of  iron  fire-boxes  is  generally  caused  by  blistering;  the  failure  of  steel 
fire-boxes  is  generally  caused  by  cracking.  But  with  careful  firing  it  has  not  been  an 
unusual  occurrence  for  a  locomotive  with  a  steel  fire-box  to  make  a  mileage  of  200,000 
miles  before  the  fire-box  failed ;  in  the  reports  of  proceedings  of  the  American  Railway 
Master-Mechanics  one  instance  is  recorded  where  a  mileage  as  high  as  505,890  miles 
was  obtained  before  the  steel  fire-box  failed. 


MODERN  LOCOMOTIl'K   CONSTRUCTION.  463 

STAY   BOLTS. 

456.  Stay  bolts  for  the  fire-box  are  usually  made  of  iron  5  inch  diameter,  with  12 
threads  per  inch ;  when  the  furnace  side  sheets  are  i  inch  thick,  the  distance  from 
center  to  center  of  stay  bolts  should  not  exceed  4  inches;  for  furnace  sheets  ^  inch 
thick  the  distance  from  center  to  center  of  stay  bolts  can  be  increased,  but  should 
never  exceed  4£  inches. 

For  high  steam  pressure  a  good  rule  is  to  space  the  stay  bolts  so  that  the  stress 
per  square  inch  of  the  smallest  cross-section  of  the  bolt  will  not  exceed  6,000  pounds 
per  square  inch.  (See  Art.  482.)  The  stay  bolts  are  screwed  into  both  sheets,  and 
riveted  over  cold. 

457.  Sometimes  one  or  two  rows  of  hollow  stay  bolts  are  placed  directly  over  the 
fuel;  the  purpose  of  these  is  to  admit  air  to  aid  combustion.     Sometimes  three  or 
four  short  pieces  of  2-inch  boiler  tubes  are  inserted  and  expanded  in  each  side  of  the 
fire-box  in  place  of  hollow  stay  bolts  for  the  same  purpose,  but  whether  such  devices 
will  accomplish  their  purpose  is  very  doubtful,  as  the  air  admitted  is  of  too  low  a 
temperature  to  ignite  the  gases. 

Occasionally  we  find  holes  drilled  into  the  two  upper  rows  of  stay  bolts  in  the 
sides  of  the  fire-box ;  these  holes  are  about  &  of  an  inch  in  diameter,  and  extend  in  the 
direction  of  the  axes  of  the  bolts  to  a  depth  of  about  $  to  1  inch  from  the  outside  of 
the  box ;  the  purpose  of  these  holes  is  to  sound  an  alarm  of  danger  in  case  any  of 
them  have  been  broken. 

458.  The  crown  sheet  is  often  stayed  in  a  similar  manner — that  is  to  say,  it  is 
stayed  by  radial  stay  bolts  screwed  into  the  outer  and   fire-box  crown  sheets,  and 
riveted  over  cold ;  these  stay  bolts  are  shown  in  Fig.  634. 

For  stays  of  this  kind  the  furnace  crown  sheet  is  of  a  necessity  curved  consider- 
ably, almost  approaching  a  cylindrical  form.  The  reason  for  this  is  to  make  the  stay 
bolts  enter  both  crown  sheets  as  nearly  as  possible  in  a  line  normal  to  both  sheets,  so 
as  to  obtain  as  many  full  threads  as  can  be  had  in  either  sheet.  Another  advantage 
gained  by  this  curvature  is  that  the  impurities  in  the  water  as  they  are  precipitated 
cannot  find  a  ready  lodgment  on  the  crown  sheet ;  the  strong  circulation  washes  the 
impurities  off  the  sheet  and  keeps  it  comparatively  clean.  Hence  this  form  of  crown 
sheet  and  the  manner  of  staying  it  are  well  adapted  for  muddy  water,  and  have  been 
favorably  received.  But  these  curved  crown  sheets  have  also  a  disadvantage,  namely, 
they  compel  us  to  throw  out  too  many  tubes  at  the  upper  corners  of  the  furnace,  and 
in  order  to  make  up  the  right  number  of  tubes  more  must  be  added  in  the  center  at 
the  top  of  furnace.  This  arrangement  compels  us  either  to  be  satisfied  with  less  than 
the  desirable  amount  of  steam  room,  or  we  must  increase  the  diameter  of  the  boiler, 
which  is  not  always  an  easy  thing  to  do. 

459.  To  overcome  the  objectionable  feature  of  this  form  of  crown  sheet  and  still 
have  the  stay  bolts  normal  to  the  sheets,  the  Belpaire  boiler,  referred  to  in  Art.  423,  has 
sometimes  been  adopted.     Besides  the  objections  to  the  Belpaire  boiler  given  in  Art. 
423,  we  may  mention  that  the  impurities  of  the  water  have  a  chance  to  settle  on  its 
furnace  crown  sheet ;  furthermore,  the  stay  bolts  through  this  sheet  tend  to  entrap  the 
impurities  of  the  water,  causing  them  to  pile  around  the  stay  bolts  in  heaps  of  conical 


464 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


form,  and  covering  a  considerable  portion  of  the  crown  sheet ;  all  this  increases  the 
liability  to  overheating,  which  may  lead  to  disastrous  results,  unless  great  care  is  taken 
to  keep  the  crown  sheet  clean  by  frequent  inspection  and  washing.  It  is  our  opinion 
that  the  Belpaire  boiler  is  not  the  right  kind  to  use  on  roads  where  pure  water  can- 
not be  obtained. 

460.  In  the  Belpaire  boilers  the  distances  between  the  centers  of  stay  bolts  in  the 
crown  sheets  vary  from  4£  to  4f  inches,  seldom  exceeding  the  latter.     In  ordinary 
boilers  with  curved  crown  sheets,  the  distances  between  the  stay  bolts  are  measured  on 
the  furnace  crown  sheet,  and  also  vary  from  4$  to  4g  inches ;  the  distances  between 
the  stay  bolts  measured  on  the  outer  crown  sheet  in  this  type  of  boilers  are  of  course 
much  greater,  but  this  is  of  no  consequence,  as  this  outer  crown  sheet  does  not  require 
as  much  bracing  as  that  in  the  Belpaire  box.     Boilers  whose  furnace  crown  sheets  are 
stayed  by  stay  bolts  usually  have  the  dome  placed  on  the  barrel.    When  the  dome  is 
placed  over  the  furnace,  crown  bars  will  have  to  be  used. 

CROWN   BARS. 

461.  Crown  bars  are  generally  made  of  two  pieces  of  bar  iron,  as  shown  in  section  7?, 
Fig.  686.     These  bars  vary  from  4  to  5J  inches  in  depth,  and  fi-om  f  to  £  inch  in  thick- 


B 


Fig*  681 


Fig.  686 


ness.  The  ends  a  are  sometimes  bent  and  welded  together,  as  shown  at  C,  Fig.  686 ; 
and  sometimes  the  ends  are  left  straight  and  a  piece  6  set  between  them  and  the  whole 
welded  together,  as  shown  at  Z),  Fig.  686.  The  remaining  portion  of  the  bars  are  left 
a  sufficient  distance  apart  for  the  crown-bar  bolts  to  pass  through. 

Sometimes  the  crown  bars  are  placed  lengthways  of  the  furnace,  but  the  usual 
practice  is  to  place  them  across  the  crown  sheet ;  by  doing  so  they  will  be  shorter,  and 
consequently,  with  the  same  cross-section,  will  be  stronger  than  if  placed  lengthways. 
The  ends  of  the  crown  bars  rest  on  the  edges  of  the  side  sheets,  and,  to  a  small  extent, 
on  the  corners  of  the  crown  sheet,  as  shown  at  a  in  Fig.  687.  The  distance  from  center 
b  of  the  crown  bar  to  center  of  the  next  one  varies  from  4J  to  5  inches,  seldom  exceed- 
ing the  latter ;  in  fact,  a  distance  of  5  inches  between  the  centers  is  somewhat  excessive ; 
and  yet  this  distance  cannot  bs  made  much  less  than  4£  inches,  as  there  must  be 
sufficient  room  left  to  put  in  the  bolts  which  connect  the  crown-bar  braces  to  the 
crown  bar.  The  distance  c  between  the  bottom  of  the  crown  bars  and  the  top  of 
crown  sheet  is  sometimes  1  inch,  but  in  the  majority  of  boilers  it  is  l£  inches ;  the 
latter  distance  we  believe  will  be  more  satisfactory,  because  it  allows  a  freer  circulation 
and  reduces  the  liability  of  the  mud  filling  up  the  spaces. 


MUHERS  LOCOMOTIVE   CONSTRUCTION. 


465 


CUOWN-BAK   BOLTS. 


462.  The  crown-bar  bolts  are  generally  J  or  1  inch  in  diameter ;  they  are  placed 
from  4J  to  5  inches  from  center  to  center,  but  the  former  distance  is  preferable.  In  many 
boilers  the  crown-bar  bolts  have  a  button-head ;  the  head  is  placed  on  the  underside  of 
the  crown  sheet ;  the  nuts  of  these  bolts  bear  against  wrought-iron  washers  which  are 


'854- 

Fig,  692 

lipped  over  the  top  of  crown  bars.  These  washers  are  made  in  a  very  easy  manner ; 
they  are  simply  square  plates  each  having  two  corners  turned  over  to  form  the  lips,  as 
shown  at  d  in  Fig.  687. 

Occasionally  the  crown-bar  bolts  have  a  T-head  extending  over  the  crown  bars ; 
the  ends  of  these  bolts  are  riveted  over  on  the  underside  of  the  crown  sheet,  as  shown 
in  Fig.  688.  Sometimes  rivets  are  used  in  place  of  bolts,  as  shown  in  Fig.  687,  and 
riveted  over  on  the  underside. 

Occasionally  the  crown-bar  bolts  are  of  a  form  as  shown  in  Fig.  689 ;  these  are 
M-rtiwed  into  the  crown  sheet,  and  a  nut  at  the  other  end  secures  the  bar  to  the  sheet. 

The  crown-bar  bolts  shown  in  Fig.  690  we  believe  to  be  the  best  kind ;  they  have 
a  slight  taper  under  the  head,  the  crown  sheet  is  reamed  out  to  fit  this  taper,  and  the 
crown  bars  are  secured  in  place  by  means  of  nuts  bearing  against  the  washers  on  the 
top.  The  advantage  of  this  form  is  that,  should  leaks  occur,  the  bolts  can  be  readily 
taken  out,  and  the  tapered  part  extended  by  turning  a  small  portion  off  the  head  and 
refitting  the  bolt  in  crown  sheet. 

Fig.  691  shows  a  bolt  of  this  kind  on  a  larger  scale,  and  Fig.  692  shows  the 
crown-bar  bolt  as  used  in  Fig.  689. 

463.  The  thimbles  between  the  crown  bars  and  crown  sheet  are  sometimes  made 
of  wrought-iron  \  or  f  inch  thick.  They  are  formed  simply  by  bending  the  iron  over 
;i  mandrel;  the  ends  are  not  welded.  Washers  of  this  kind  cover  up  too  much  of  the 
crown  sheet.  To  overcome  this  objection  the  washers  are  often  made  of  cast-iron  and 
tapered  towards  the  crown  sheet,  as  shown  in  Fig.  688,  leaving  at  the  bottom  of  the 


466  MODERN  LOCOMOT1VK   CONSTRUCTION; 

washer  about  %  of  an  inch  metal  around  the  bolt  ;  the  upper  diameter  is  made  suffi- 
ciently large  to  cover,  or  nearly  so,  the  whole  width  of  the  crown  bar.  With  these 
washers  a  greater  surface  of  the  crown  sheet  is  in  contact  with  the  water,  and  over- 
heating around  the  bolt  is  not  so  liable  to  occur. 

CROWN-BAK  BRACES. 

464.  For  ordinary  locomotives  there  should  be  at  least  four  braces  running  from 
each  crown  bar  to  the  outer  crown  sheet.  The  appearance  of  so  many  crown  bars  on 
the  top  of  furnace  may  give  an  impression  that  the  braces  are  superfluous,  but  such  is 
not  the  case,  as  can  be  easily  shown  by  computing  the  load  which  a  crown  bar  can 
support. 

For  example  :  Say  the  crown  bar  is  54  inches  long  —  this  length  we  may  assume  to 
be  the  distance  from  inside  to  inside  of  the  side  sheets  on  which  the  ends  of  the  crown 
bar  rest  ;  let  the  depth  be  5  inches  ;  and  the  thickness,  f  inch  ;  it  is  required  to  find 
the  load  which  the  crown  bar  can  support  without  being  reinforced  by  any  braces. 

For  all  practical  purposes  we  may  assume  the  crown  bar  to  be  a  beam  supported 
at  the  ends  with  a  load  uniformly  distributed. 

The  load  which  any  bar  of  a  uniform  rectangular  cross-section  can  support  is 
found  by  the  following  formula  : 


61 

in  which  W  denotes  the  load  in  pounds;  S,  the  safe  stress  per  square  inch  on 
the  fibers  most  remote  from  the  neutral  surface;  for  wrought-iron  this  stress  is 
usually  taken  at  12,000  pounds  ;  b,  the  breadth  ;  d,  the  depth  ;  and  /  the  length,  all  in 
inches.  The  value  of  n  will  depend  on  the  manner  in  which  the  beam  is  supported 
and  loaded  ;  in  this  case  the  value  of  n  is  8.  Hence  the  foregoing  formula  may  be 
written  as  follows  : 


6  x  I 
This  formula  may  again  be  reduced  to 

_  16000  x  b  x  d2 
I 

Hence  from  the  foregoing  we  have  the  following  rule:  Multiply  16,000  by  the 
breadth,  and  by  the  square  of  the  depth,  all  in  inches,  and  divide  the  product  by  the 
length  in  inches ;  the  quotient  will  be  the  load,  uniformly  distributed,  which  the  crown 
bar  can  safely  support. 

We  have  seen  that  the  crown  bar  is  made  of  two  pieces,  each  of  which  is  f  inch  in 
thickness ;  we  may  therefore  say  that  the  total  thickness  of  the  crown  bar  is  l£  inches. 
Now,  substituting  for  the  symbols  in  the  last  formula  their  values,  we  have 

16000  x  1.5  x  52 

— ^r~         ~  =  11,111  pounds  (the  fraction  being  neglected), 

and  this  is  the  load,  uniformly  distributed,  which  the  crown  bar  can  support. 


MODERN  LOCOMOTirK   CONSTRUCTION.  4(J7 

Now  let  us  compute  the  load  which  the  crown  bar  has  to  support.  Assume  that 
these  bars  are  placed  5  inches  from  center  to  center,  and  that  the  steam  pressure  is 
150  pounds  per  square  inch.  The  total  steam  pressure  on  a  strip  of  the  crown  sheet 
5  inches  wide  and  54  inches  long  will  be 

5  x  54  x  150  =  40500  pounds. 

Now,  neglecting  the  load  which  the  sheet  itself  can  support,  we  may  say  that  this 
is  the  load,  uniformly  distributed,  which  the  crown  bar  must  support  under  the  given 
condition,s. 

But  we  have  found  by  computation  that  the  crown  bar  can  support  only  11,111 
pounds,  which  leaves  40500  —  11,111  =  29389  pounds  which  must  be  supported  by 
the  braces. 

This  result  also  enables  us  to  compute  the  cross-sectional  area  of  the  braces.  These 
should  never  be  subjected  to  a  greater  tensile  stress  than  6,000  pounds  per  square  inch 
of  cross-section.  Now,  adopting  6,000  pounds  for  the  stress  per  square  inch,  and 
dividing  the  total  load  which  the  braces  have  to  support  (29,389  pounds)  by  this  stress 
per  square  inch,  we  obtain  the  aggregate  number  of  square  inches  in  the  cross- 
sectional  area  of  the  braces,  thus  : 

29389 

=  4.89  square  inches. 


We  may  assume  that  each  end  of  the  crown  bar  sustains  one-half  the  load  of  one 
brace,  and  if  there  are  to  be  four  braces  to  each  crown  bar,  and  we  wish  to  find  the 
cross-sectional  area  of  one  brace,  we  proceed  as  if  there  are  five  braces  to  sustain  a 
load  of  29,389  pounds.  Now,  dividing  the  total  cross-sectional  area  of  the  braces 
as  previously  found,  namely,  4.89  square  inches,  by  the  number  of  braces,  we  have 

4.89 

-  =  0.97  square  inch  for  each  brace  ;  the  corresponding  diameter  is  nearly  1  J  inches. 

In  a  similar  way  we  find  the  diameters  of  the  braces  for  any  other  steam  pressure. 

TRANSVERSE  BRACES. 

465.  The  transverse  braces  b,  which  are  placed  above  the  furnace  crown  sheet, 
Fig.  693,  for  staying  the  outer  sides  of  the  fire-box,  are  secured  to  the  side  sheets  in 
different  ways.  Sometimes  they  are  screwed  into  the  side  sheets  and  then  riveted 
over  cold,  as  shown  at  o,  Fig.  694.  In  cases  of  this  kind  one  transverse  brace  b  is 
placed  in  each  space  between  the  crown  bars  (see  Fig.  693).  Then,  again,  we  find  the 
transverse  braces  fastened  to  crow-feet  c  by  means  of  pin  bolts,  Fig.  695.  In  cases  of 
this  kind  a  transverse  brace  is  placed  in  every  other  space  between  the  crown  bars, 
as  shown  in  Fig.  696,  or  in  every  space  between  the  crown  bars,  as  shown  in  Fig. 
697.  When  these  braces  are  placed  in  every  other  space  the  crow-  feet  c  stand 
horizontally,  as  shown  in  Fig.  696,  and  when  placed  in  every  space  the  crow-feet  c 
stand  vertically,  as  shown  in  Fig.  697.  The  crow-feet  are  riveted  to  the  side  sheets  ; 
each  crow-foot  has  two  rivets.  These  rivets  are  laid  off  in  straight  lines  generally 
parallel  to  the  furnace  crown  sheet  •  the  horizontal  distance  between  the  centers  of 


468 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


these  rivets  will  of  course  depend  on  the  distance  between  the  centers  of  crown  bars ; 
if  the  latter  are  spaced  4£  inches  apart,  the  horizontal  distance  between  the  centers  of 
rivets  will  also  be  4 £  inches ;  furthermore,  these  rivets  will  genet-ally  fall  in  the  ver- 


Fig.  695 


Pig.  696 


tical  lines  drawn  through  the  centers  of  stay  bolts,  or  nearly  so,  giving  the  rivets  the 
appearance  of  rows  of  stay  bolts. 

466.  It  may  be  well  here  to  draw  attention  to  the  horizontal  seams  d  of  the  outer 
sides  and  crown  sheets,  Fig.  698.      Considerable  ti'ouble  will  be  experienced  in  the 

attempt  to  place  one  of  these  seams 
between  two  rows  of  stay  bolts  or  be- 
tween the  upper  row  of  stay  bolts  and 
crow-feet,  on  account  of  insufficient 
room.  Consequently  these  seams  are 
generally  placed  above  the  crow-feet. 
Sometimes  the  side  sheets  and  outer 
crown  are  all  made  of  one  sheet,  thus 
avoiding  the  side  seams.  This  we  be- 
Fig.  697  lieve  to  be  the  best  practice,  although 

there  is  one  objection  to  it,  namely, 
it  makes  the  sheet  very  heavy  for  handling,  and  to  flange  the  dome  opening  is  some- 
what troublesome,  unless  the  boiler  shop  is  specially  fitted  up  for  the  purpose. 


BOILER  BRACES. 

467.  Considerable  space  is  left  between  the  bottom  of  the  tubes  and  the  upper 
row  of  stay  bolts  in  the  furnace  tube-sheet  (see  Fig.  624),  and  this  portion  of  the  tube 
sheet  must  be  stayed.  For  this  purpose  braces  d,  Fig.  623,  are  run  from  the  tube 
sheet  to  the  bottom  of  the  shell.  These  braces  are  sometimes  attached  to  the  crow- 
feet which  are  riveted  to  the  tube  sheet,  and  the  other  ends  of  the  braces  are  riveted 
directly  to  the  shell.  Since  this  portion  of  the  tube  sheet  is  exposed  to  a  great  heat, 
it  is  advisable  to  place  cast-iron  thimbles  of  the  same  form  as  those  under  the  crown 
bars  between  the  tube  sheet  and  the  crow-feet,  so  as  to  allow  as  great  a  surface  as 
possible  of  the  former  to  come  in  contact  with  the  water,  thereby  preventing  over- 
heating. 

The  tapered  course  J?,  Fig.  623,  leaves  another  flat  place  in  each  of  the  sides  of  the 
shell  which  must  be  stayed,  Sometimes  they  are  stayed  by  braces  running  across  the 


MODERN  LOCOMOTIVE  CONSTRUCTION.  4(50 

boiler;  in  such  cases,  of  course,  the  tubes  are  spaced  so  as  to  leave  room  for  these 
braces.  But  the  general  practice  is  to  rivet  stiffening  pieces  to  the  sides,  as  shown  at 
/  in  Fig.  623.  These  stiffening  pieces  are  often  made  of  bar  iron  about  2J  inches 
square,  and  we  believe  that  this  is  the  best  form.  We  recommend  a  distance  of  3J 
inches  from  center  to  center  of  these  pieces.  Their  length  should  be  such  as  to 
extend  a  foot  or  so  both  ways  on  the  cylindrical  part  of  the  shell.  Sometimes  the  stiff- 
ening pieces  f  are  made  of  T-iron,  as  shown  in  Fig.  655,  and  then  again  we  find  them 
made  of  angle  iron  placed  back  to  back. 

The  manner  of  staying  the  back  head  is  plainly  shown  in  Figs.  624,  634,  and  657. 
Crow-feet  are  generally  riveted  to  the  back  head,  and  braces  g  (Fig.  655)  extend  from 
these  to  the  shell  of  the  boiler.  Sometimes  T-irons  are  used  in  place  of  crow-feet. 

The  front  tube  sheet  is  stayed  in  a  manner  similar  to  that  of  the  back  head.  In 
all  cases  the  rivets  which  secure  the  crow-feet  or  T-iron  to  the  back  head  and  front 
tube  sheet  should  be  spaced  so  as  to  bring  them  4£  inches  apart  from  center  to  center, 
and  arranged  in  rows  like  the  stay  bolts.  The  braces  should  be  placed  as  nearly  hori- 
zontal as  possible,  so  as  to  reduce  the  stress  in  them.  Again,  the  ends  of  these  braces 
which  are  riveted  to  the  shell  should  be  kept  a  considerable  distance  apart,  so  as  to  dis- 
tribute the  stress  over  the  shell  as  much  as  possible.  For  computing  the  stress  in  these 
braces  see  Art.  485. 

The  braces  g  which  we  just  described  are  usually  called  diagonal  braces,  to  dis- 
tinguish them  from  tie  rods,  which  are  sometimes  used  in  place  of  the  diagonal  braces. 

468.  The  tie  rod  is  shown  in  Fig.  699.  It  extends  from  the  back  head  to  the  front 
tube  sheet,  and  it  has  a  hex- 
agonal head  at  one  of  its  ends. 
The  diameter  of  the  portion 
near  the  head  is  larger  than 
that  of  the  body  of  the  rod ; 
this  enlarged  part  is  about 
1  inch  long,  with  a  thread 
cut  on  it,  and  is  screwed  into  the  back  head.  In  order  to  secure  a  steam-tight  joint  a 
copper  washer  is  placed  between  the  head  and  the  sheet. 

The  end  b  of  the  rod  is  also  made  larger  in  diameter  than  that  of  the  body  of  the 
rod,  and  has  a  thread  cut  on  it,  but  it  is  not  so  large  in  diameter  as  the  part  near  the 
head.  If,  for  instance,  the  body  of  the  rod  is  5  inch  in  diameter,  the  part  near  the  head 
will  be  H  inches  diameter  outside  of  the  thread,  and  the  end  b  will  be  1  inch  diameter, 
also  measured  outside  of  the  thread.  The  end  b  of  the  rod  is  not  screwed  into  the  front 
tube  sheet,  but  simply  passed  through  it,  and  is  secured  to  the  tube  sheet  by  means  of 
two  nuts,  one  inside  and  the  other  outside  of  the  sheet;  copper  washers  are  placed 
between  the  sheet  and  the  nuts.  Great  care  must  be  taken  in  screwing  up  the  inner 
nut,  which  should  bear  but  slightly  against  the  sheet,  so  as  to  avoid  buckling  the  rod ; 
the  joint  is  made  steam-tight  by  screwing  up  tightly  the  outer  nut. 

These  tie  rods  are  generally  placed  5  or  5£  inches  from  center  to  center.  The 
diameter  of  the  rods  will  depend  on  the  steam  pressure,  and  should  be  such  that  the 
stress  per  square  inch  will  not  exceed  6,000  pounds.  To  illustrate  we  will  take  the 
following  example : 


470  MODERN  LOCOMOTIVE   CONSTRUCTION, 

EXAMPLE  143.  —  What  should  be  the  diameter  of  the  tie  rods  placed  5  inches  from 
center  to  center,  steam  pressure  150  pounds  per  square  inch  ? 

The  total  surface  which  each  tie  rod  has  to  support  is  5  x  5  =  25  square  inches  ; 
and  since  the  steam  pressure  is  150  pounds  per  square  inch,  the  total  tension  on  the 
rod  will  be  25  x  150  '=  3750  pounds.  The  stress  is  not  to  exceed  6,000  pounds  per 
square  inch,  hence  the  cross-sectional  area  of  the  smallest  part  of  the  rod  will  be 

3750 

—  -625  square  inch  ;  the  corresponding  diameter  is  nearly  ff  inch. 


469.  Fig.  647  shows  a  boiler  in  which  the  back  head  is  stayed  by  gusset  plates 
instead  of  diagonal  braces  or  tie  rods.     Sometimes  the  front  head  is  stayed  in  a  similar 
way.     The  usual  way  of  fastening  the  gusset  plates  to  the  boiler  sheets  is  to  rivet 
their  ends  to  two  angle  irons,  and  these  are  in  turn  riveted  to  the  boiler  sheets. 

DOMES. 

470.  A  common  way  of  fastening  the  dome  to  the  shell  is  shown  in  Fig.  655.    As 
will  be  seen,  the  outer  crown  is  flanged  up  into  the  dome,  and  the  dome  is  flanged  to 
fit  the  outer  crown. 

The  dome  flange  is  riveted  to  the  shell  by  a  single  row  of  rivets  ;  their  pitch  is  to 
be  taken  from  the  tables  in  Art.  449.  The  vertical  flange  of  the  shell  is  also  riveted  to 
the  dome  by  a  single  row  of  rivets,  but  the  pitch  of  these  is  double  that  of  the  rivets 
through  dome  flange  and  outer  crown.  It  is  often  assumed  that  this  manner  of  fasten- 
ing the  dome  to  the  shell  is  sufficiently  strong,  but  our  experience  has  led  us  to  believe 
that  it  is  not  strong  enough,  and  is  particularly  weak  for  the  high  steam  pressure 
adopted  in  late  years.  Every  dome  should  have  a  wrought-iron  stiffening-ring  riveted 
to  the  inside  of  boiler,  as  shown  in  Fig.  633  or  Fig.  647.  The  rivets  through  the  dome 
flange  and  shell  should  also  pass  through  this  ring,  and  in  addition  to  these  rivets 
another  row  should  pass  through  the  ring  and  shell  ;  the  pitch  of  the  rivets  in  this  last 
row  is  usually  double  that  of  the  rivets  through  the  dome  flange.  Without  this  stiff- 
ening ring  the  boiler  is  liable  to  spread  directly  under  the  dome,  thereby  throwing  an 
excessive  stress  on  the  dome  flanges,  and  causing  leaks.  This  action  has  been  fully 
confirmed  by  experiments  on  the  New  York  Elevated  Railroad. 

SHOET   SMOKE-BOXES. 

471.  A  few  years  ago,  short  smoke-boxes,  shown  in  Figs.  700  and  701,  were  almost 
universally  used,  and  even  at  the  present  time  they  are  often  put  on.     The  length  of 
these  boxes  is  such  as  to  give  but  very  little  more  than  sufficient  room  for  the  steam 
pipes,  exhaust  nozzles,  and  draft  or  petticoat  pipe. 

The  exhaust  nozzle  (not  shown  here)  must  necessarily  be  short.  The  petticoat 
pipe  a,  through  which  the  exhaust  steam  passes,  is  made  telescopic,  so  that  it  can  be 
lengthened  or  shortened  ;  the  purpose  of  changing  its  length  is  to  obtain  an  even  draft 
on  the  fire.  This  pipe,  after  it  has  been  correctly  adjusted  in  position,  is  held  by 
the  braces  b  which  extend  to  the  sides  of  the  smoke-box,  one  on  each  side  of  the 
pipe.  Boxes  of  this  kind  have  smoke-stacks  of  the  form  as  shown  in  Figs.  769,  770,  of 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


471 


which  a  description  is  given  further  on.  These  boxes  are  of  no  use  for  holding 
sparks ;  they  are  not  designed  for  such  a  duty,  consequently  the  sparks  are  thrown  out 
of  the  stack  while  the  engine  is  running,  thereby  frequently  doing  much  harm  by  set- 


\\~oxoxoxo 
•l\  OXOXOXQ 

x 


Fig.  703 


ting  fire  to  objects  along  the  road,  and  putting  the  railroad  companies  to   a  great 
expense  in  paying  for  the  damages. 

Figs.  702  and  703  represent  another  smoke-box,  which  may  be  properly  classed 
with  the  short  ones.  It  is  used  on  the  New  York  Elevated  Railroad  engines,  and  is 
designed  for  a  straight  stack  similar  to  the  one  shown  in  Fig.  778  or  Fig.  781.  Stacks 


472 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


of  this  kind  have,  of  course,  no  netting,  consequently  the  netting  is  placed  inside  of  the 
smoke-box.  The  exhaust  nozzles  extend  above  the  horizontal  netting  A  B,  and  this 
netting  terminates  at,  and  is  fastened  to,  the  diaphragm  plate  D  E.  The  lower 
portion  of  this  plate  is  adjustable,  and  answers  the  same  purpose  as  the  petticoat  pipe, 
namely,  to  create  an  even  draft  on  the  fire.  In  this  box  there  is  some  room  provided 
for  retaining  the  sparks,  and  for  short  runs  it  is  well  suited. 


EXTENSION  FRONTS. — BRICK  ARCHES. 


472.  Various  spark  arresters  have  been  invented  and  designed,  but  when  put 
into  service  proved  themselves  to  be  of  little  or  of  no  practical  value. 

The  extended  smoke-boxes,  or  extension  fronts,  as  they  are  sometimes  called, 
shown  in  Figs.  704  and  705,  were  condemned  after  they  were  first  brought  out,  but 


Fig.  705 


they  were  tried  again,  and  in  late  years  have  been  received  with  great  favor,  and  we 
believe  justly  so,  for  they  certainly  retain  a  large  quantity  of  the  sparks,  and  to  a 
great  extent  prevent  fires  along  the  road.  The  extension  front  has  its  enemies,  but  it 
has  more  friends,  and  is  considered  by  many  railroad  men  to  be  one  of  the  greatest 
boiler  improvements  made  in  late  years.  Comparing  Fig.  704  with  Fig.  701,  we  find  that 
the  boxes  are  practically  alike,  with  the  exception  that  the  extension  front  is  about 
twice  as  long  or  a  little  more  than  the  short  box.  The  primary  object  of  the  extension 
front  is  to  hold  the  sparks  and  carry  them  to  a  designated  place  where  the  box  can  be 
emptied  and  the  sparks  can  do  no  harm,  instead  of  scattering  them  along  the  road  to 
stai't  fires,  the  worst  use  that  sparks  can  be  put  to.  But  besides  this  advantage  which 
the  extension  front  possesses,  it  is  also  claimed  that  it  will  save  fuel.  This  may  be 
explained  as  follows :  "We  have  already  seen  that  with  a  short  box  the  sparks  are 
thrown  out  of  the  stack,  and  we  may  here  add  that  many  of  the  sparks  and  cinders 
are  broken  up  in  the  stack  by  being  thrown  violently  against  the  netting.  Now  to 


MOVERS  LOCOMOTIVE   CONSTRUCTION.  473 

lift  these  sparks  and  cinders,  and  give  them  the  velocity  which  they  attain  in  the  stack, 
Invak  up  some  of  them,  and  throw  all  out  of  the  stack,  requires  an  expenditure  of 
energy,  and  this  energy  is  supplied  by  the  exhaust  steam.  Besides  this,  the  exhaust 
steam  has  its  legitimate  work  to  do,  namely,  to  create  a  forced  draft.  In  the  extension 
front  some  of  the  dust  and  finer  cinders  will  also  be  thrown  out  of  the  stack,  and 
indeed  this  cannot  be  avoided  so  long  as  an  opening  for  the  escape  of  the  exhaust  is 
required.  But  it  is  also  a  fact  that  the  extension  front  becomes  filled  with  cinders, 
and  therefore  not  near  so  many  cinders  are  thrown  out  of  the  stack,  neither  are  these 
broken  up  as  in  the  short  box,  consequently  the  exhaust  steam  in  the  extension  front 
has  less  work  to  do  than  in  the  short  box.  Again,  in  the  extension  front  the  exhaust 
nozzle  A  (Fig.  704)  extends  to  above  the  netting  E,  and  the  exhaust  steam  passes  out 
through  a  straight  stack,  and  meets  no  obstruction ;  this  greater  freedom  of  escape 
also  reduces  the  work  of  the  exhaust  steam.  We  may  therefore  say  that  nearly  the 
whole  force  of  the  exhaust  steam  is  expended  in  creating  the  required  draft,  and  it  has 
no  extra  work  to  perform.  Now,  since  in  the  extension  front  the  exhaust  steam  has 
less  work  to  perform  than  in  the  short  box,  we  can  make  the  exhaust  nozzles  a  little 
larger  in  diameter,  and  in  fact  they  are  usually  made  larger  in  an  extension  front  than 
in  a  short  box.  But  increasing  the  diameter  of  the  exhaust  nozzle  will  give  less  back 
pressure  in  the  cylinder ;  it  also  gives  a  smoother  draft,  and  is  not  so  liable  to  draw 
unconsumed  fine  coal  through  the  flues.  Hence  a  lighter  fire  may  be  carried,  and,  in 
fact,  lighter  fires  are  generally  carried,  in  locomotives  with  an  extension  front  than  in 
others ;  this  also  gives  more  time  for  the  fuel  to  burn.  All  these  things  tend  to  and  do 
save  fuel,  but  there  is  such  a  thing  as  overestimating  the  advantages  of  an  extension 
front. 

In  many  engines  a  brick  arch,  as  is  shown  in  Fig.  668,  is  used  in  connection  with 
the  extension  front,  and  since  this  arch  prevents  unconsumed  fuel  from  being  drawn 
through  the  flues,  it  must  also  promote  the  economy  of  fuel.  The  exact  amount  of 
fuel  saved  by  the  brick  arch  and  that  saved  by  the  extension  front  cannot  be  stated, 
as  reliable  experiments  to  determine  these  amounts  have  not,  to  our  knowledge,  been 
made.  But  that  the  brick  arch  in  connection  with  the  extension  front  does  save  fuel 
is  beyond  a  doubt. 

473.  The  comparatively  large  capacity  of  the  extension  front  affects  to  a  slight 
extent  the  useful  action  of  the  exhaust,  because  the  exhaust  steam  has  to  create  a 
partial  vacuum  in  a  larger  space  than  in  a  short  box,  and  this  increased  capacity 
absorbs  a  part  of  the  exhausting  action  of  the  blast  before  the  tubes  are  appreciably 
affected ;  and  it  may  be  said  that,  within  limits,  the  smaller  we  make  the  capacity  of  the 
smoke-box,  the  more  will  the  exhausting  action  be  felt  by  the  tubes.    Mr.  D.  K.  Clark, 
in  his  treatise  on  Railway  Machinery,  gives  for  the  most  suitable  capacity  of  the  smoke- 
box  3  cubic  feet  for  every  square  foot  of  grate.     Now  the  capacity  of  the  modern 
extension  front  is  not  much  larger  than  this,  and  when  partly  filled  with  cinders  is  in 
some  cases  less.     We  may  therefore  conclude  that  although   the  increased  capacity 
may  affect  the  useful  action  of  the  exhaust,  it  does  not  affect  it  to  any  hurtful  extent. 

474.  That  the  extension  front  has  sometimes  been  a  failure  cannot  be  denied, 
but  this  was  probably  due  to  faulty  design,  bad  workmanship,  and  lack  of  attention 
when  the  engine  was  running.     In  order  to  make  an  extension  front  work  successfully 


474 


MOllKHX 


COXSTRVCTION. 


it  must  be  fitted  up  perfectly  air-tight,  otherwise  the  smoldering  cinders  which  it 
holds  will  take  fire,  which  will  warp  the  sheets  and  render  the  box  useless.  Care 
should  also  be  taken  not  to  allow  the  front  to  fill  up  too  much ;  such  a  condition  will 
cause  the  sparks  to  be  thrown  out  of  the  stack.  The  diaphragm  plate  C  (Fig.  704)  and 
netting  E  must  be  properly  arranged,  also  suitable  exhaust  nozzles  adopted  and 
ample  provision  made  for  emptying  the  box. 

The  diaphragm  plate  G  is  generally  made  solid ;  sometimes  its  lower  part  D  is 
perforated,  as  shown  in  Fig.  705.  This  plate  C  is  fastened  to  the  tube  sheet  by  an 
angle  iron  placed  directly  over  the  top  row  of  tubes. 

The  horizontal  distance  a  from  the  lower  edge  of  the  diaphragm  to  the  tube  sheet 
is  greater  than  the  distance  b  between  the  plate  and  tube  sheet  at  the  top.  The 
distance  a  at  the  bottom  will  depend  on  the  number  of  tubes;  it  should  be  such  as  to 
give  a  horizontal  area  for  the  passage  of  gases  equal  to  the  total  cross-sectional  area 
of  the  tubes.  At  the  top  a  distance  of  2£  to  3  inches  is  usually  sufficient.  We  have 
already  seen  that  the  lower  part  of  the  diaphragm  plate  is  adjustable,  or,  in  other 
words,  it  is  made  so  that  it  can  be  raised  or  lowered,  and  the  extent  of  the  total 
adjustment  is  about  6  inches.  The  means  provided  for  fastening  the  lower  portion  D 
of  the  diaphram  to  the  upper  part  consist  generally  of  four  £-inch  bolts  passed 
through  the  upper  plate  and  sliding  in  suitable  slots  cut  into  the  lower  plate,  as 
shown  in  Fig.  705.  When  the  lower  portion  of  the  diaphragm  plate  is  in  its  lowest 
position,  the  area  in  a  vertical  plane  between  its  bottom  edge  and  the  shell  of  the 
smoke-box  should  not  be  less  than  the  total  cross-sectional  area 
of  the  tubes. 

A  horizontal  plate  B  (Fig.  704)  is  usually  fastened  to  the 
upper  part  of  the  diaphragm  plate ;  it  extends  a  little  beyond 
the  exhaust  nozzle  towards  the  front  of  the  box ;  the  purpose  of 
this  is  to  prevent  the  sparks  from  flying  into  the  smoke-stack. 
The  netting  is  bolted  to  this  plate  and  extends  to  the  top  of 
smoke-box  as  shown.  The  netting  must  be  provided  with  a  door 
so  as  to  give  ready  access  to  the  upper  part  of  the  steam  pipes 
and  exhaust  nozzles. 

A  hole  F  (Fig.  704)  is  cut  in  each  side  of  the  smoke-box 
to  afford  means  for  looking  into  it  and  examining  its  condition. 
These  holes  are  closed  by  cast-iron  caps  which  fit  cast-iron 
flanges  riveted  to  the  shell.  Of  course  care  must  be  taken  to 
make  an  air-tight  joint  between  the  caps  and  flanges.  Different 
views  of  these  caps  and  flanges  are  shown  in  Figs.  706,  707, 
and  708. 

475.  The  cast-iron  cinder-box  G  for  discharging  the  cin- 
ders, shown  at  the  bottom  of  the  smoke-box,  Figs.  704  and  705, 


Kgi  706 


Fig.  708 


is  usually  riveted  to  the  shell,  and  is  generally  closed  by  means  of  a  cast-iron 
wedge  fitted  air-tight.  Sometimes  means  are  provided  for  blowing  out  the  cinders 
by  steam.  One  form  of  cinder-box  is  shown  on  a  larger  scale  in  Figs.  709,  710.  The 
flange  G  is  riveted  to  the  smoke-box.  The  opening  in  this  flange  is  closed  by  the 
wedge  H,  which  is  driven  home  air-tight  in  the  passage-way  formed  between  the 


MODERN  LOCOMOTIVK   CONSTRUCTION. 


475 


flanges  G  and  /.  A  wrought-irou  pipe  J  is  riveted  to  the  flange  I  for  the  purpose  of 
preventing  the  cinders  and  ashes  from  falling  on  the  truck  when  the  snioke-box  is 
to  be  emptied.  The  flanges  G  and  I,  also  the  wedge  //,  are  planed  tapering,  and  the 
wedge  //  fits  the  planed  surface  of  the  passage-way  very  accurately,  so  as  to  make 
an  air-tight  fit.  The  cinder-box  in  Fig.  704  is  shown  in  Figs. 
711,  712,  and  713,  on  a  larger  scale,  with  the  attachments  for 
opening  and  closing  it.  The  construction  of  this  box  is  similar 
to  that  shown  in  Fig.  710,  with  this  difference,  that  the  wedge  H, 


J-I-S5- *| 

/  -i°? 


Fig.  710 


instead  of  being  driven  home,  is  pulled  out  of  or  forced  in  the  passage-way  by  means  of 
the  screw  A'.  The  hand-wheel  L  on  this  screw  is  placed  outside  of  the  engine  frame 
M.  This  makes  a  very  convenient  arrangement  for  opening  and  closing  the  cinder-box. 

Cinder-boxes  are  sometimes  made  much  deeper  than  those  shown  in  any  of  the 
figures.  The  objection  to  deep  cast-iron  cinder-boxes  is  that,  should  the  cinders  in 
them  become  ignited  through  a  small  leakage  of  air,  the  box  is  liable  to  become  red- 
hot,  and  in  this  condition  crack  or  break  off,  thereby  rendering  the  whole  extension 
front  useless. 

476.  The  shell  of  the  smoke-box  is  sometimes  made  in  two  courses ;  this  is  done 
for  convenience  of  handling  the  plates.  These  courses  butt  against  each  other,  and 
are  connected  by  two  heavy  wrought-iron  rings  0,  P,  Fig.  704,  one  riveted  to  each 
course ;  the  rings  are  then  bolted  together  as  shown.  The  length  of  the  courses  should 
be  such  as  to  enable  the  ring  0  in  the  inner  course  to  take  the  bolts  through  the 
cylinder  saddle  and  smoke-box.  We  prefer  to  make  the  smoke-box  of  one  sheet,  even 
if  it  is  somewhat  heavy  to  handle.  In  this  case  we  also  need  a  heavy  wrought-irou 
ring  0,  as  shown  in  Fig.  714,  placed  in  a  position  to  take  the  bolts  through  the  cylin- 
der saddle.  The  purpose  of  this  ring  is  to  stiffen  the  smoke-box  and  distribute  the 
stress  through  a  larger  part  of  it. 


476 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


SMOKE-BOX   FRONTS. 


477.  The  manner  of  closing  up  the  front  end  of  the  smoke-box  is  plainly  shown  in 
Fig.  704.  An  angle-iron  ring  is  riveted  to  the  outer  end  of  the  box,  to  which  the 
cast-iron  front  is  bolted.  Frequently  a  wrought-iron  ring  3  x  l£  inches  in  cross- 
section  is  used  in  place  of  the  angle-iron  ring.  The  front  consists  of  a  cast-iron  ring 


Fig.  715 


E  and  a  cast-iron  door  S  sufficiently  large  in  diameter  to  permit  the  taking  out  of 
any  one  of  the  tubes.  The  door  is  hinged  to  the  ring,  and  in  closing  it  is  made  air- 
tight by  means  of  red-lead,  and  is  then  fastened  by  six  or  eight  clamps  T  attached  to 
the  outside  of  the  ring  K.  Sometimes  the  clamps  are  placed  inside ;  in  many  such 
cases  there  are  only  three  or  four  clamps  used,  and  arranged  so  as  to  unlock  simul- 
taneously by  one  handle  placed  outside.  We  "believe  the  plan  of  placing  the  clamps 
outside,  as  shown  in  Fig.  704,  is  the  best  one,  as  by  this  method  a  tighter  joint  can 
be  secured. 

The  cast-iron  ring  and  door  should  be  made  sufficiently  heavy  and  of  such  forms 
as  are  best  adapted  to  prevent  them  from  warping.  The  forms  shown  in  Fig.  704  we 
believe  to  be  good  ones.  The  door  is  provided  in  many  cases  with  a  wrought-iron 
liner  about  -fg  inch  thick,  placed  a  short  distance  from  the  inside  of  the  door.  Some- 
times the  liner  is  made  of  cast-iron,  as  shown  at  u,  Fig.  704.  The  object  of  this  liner 
is  to  protect  the  door  as  much  as  possible  from  excessive  heat. 

478.  In  the  extension  front  shown  in  Fig.  704,  a  petticoat  pipe  Q  is  introduced, 
but  this  is  not  a  general  practice ;  in  the  majority  of  extension  fronts  a  pipe  of  this 
kind  is  not  used. 

Long  exhaust  pipes  are  always  adopted  for  extension  fronts,  bringing  the  top  of 
the  exhaust  nozzles  to  within  a  short  distance  from  the  top  of  box. 

In  Art.  471  we  have  seen  that  in  short  smoke-boxes  the  exhaust  nozzle  is  placed 
comparatively  low,  and  the  exhaust  steam  is  led  through  the  petticoat  pipe  into  the 
stack.  The  following  reason  for  the  difference  in  the  lengths  of  the  exhaust  nozzles 


Mitl>KIt\  LOCOMOTTTS   CONSTRUCTION. 


477 


iu  the  two  kinds  of  smoke-boxos  may  be  given.  In  the  short  box  shown  in  Fig.  701 
diaphragm  plates  are  not  used,  and  therefore  an  adjustable  petticoat  pipe  is  needed 
to  make  the  draft  pull  evenly  all  over  the  fire,  and  consequently  the  exhaust  noz- 
zles must  be  placed  low,  so  that  the  object  for  which  the  petticoat  pipe  has  been 
designed  can  be  accomplished.  On  the  other  hand,  in  the  extension  front  the  pull  of 
the  draft  on  the  fire  is  regulated  by  the  lower  portion  of  the  diaphragm  plate,  and 
there  is  nothing  to  prevent  the  exhaust  nozzle  from  being  placed  pretty  close  to  the 
top  of  the  box  and  thus  leading  the  exhaust  steam  directly  into  the  stack  without 
giving  it  an  opportunity  to  spread  in  the  smoke-box. 


BRICK   ARCHES. 

479.  In  Fig.  668  the  ordinary  form  of  a  brick  arch  for  furnaces  is  represented. 
It  is  usually  4  inches  thick,  and  is  supported  by  four  wrought-iron  tubes.  These 
tubes  are  generally  1|  inches  outside  diameter  and  about  J  inch  thick — in  fact,  they  are 
made  of  the  same  kind  of  pipe  as  is  used  for  water  grates  (see  Art.  493).  The  two 
central  supporting  tubes,  as  shown  in  this  particular  case,  extend  from  the  tube  sheet 
to  the  door  sheet;  the  two  outer  ones  extend  to  the  crown  sheet.  The  brick  arch 
runs  through  the  whole  length  of  furnace  and  touches  the  sides,  excepting  at  the  two 
rear  corners  of  the  furnace,  where  openings  about  12  inches  square  are  left  for  the 
gases  to  pass  through. 

Sometimes  the  brick  arch  extends  only  through  a  part  of  the  furnace,  as  shown 
in  Fig.  716.  The  arch  shown  in  this  figure  is  a  hollow  one.  Air  is  admitted  into 


the  space  between  the  bricks  through  openings  A  in  the  front  end  of  the  fire-box ; 
these  openings  are  made  by  placing  short  pieces  of  2-inch  tube  through  the  front 
water  space,  and  expanding  the  ends  of  those  tubes  in  the  front  leg  and  tube  sheet. 
The  object  of  the  hollow  brick  arch  is  to  raise  the  temperature  of  the  air  np  to  or 
over  400  degrees  before  it  is  allowed  to  mingle  with  the  gases.  This  heated  air  is 


478 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


then  permitted  to  escape  through  the  openings  at  the  other  end  of  the  arch,  and  mingle 
with  the  gases  above  the  arch ;  the  air,  being  thus  raised  to  a  high  temperature,  will 
ignite  the  gases  more  readily  than  at  a  lower  temperature,  and  thus  produce  a  more 
perfect  combustion.  Hollow  brick  arches  require  more  care  and  give  more  trouble 
than  solid  ones,  and  therefore  the  former  are  not  so  often  used.  In  a  few  cases  the 
arch  is  supported  by  angle  irons  fastened  to  the  side  sheets  of  the  furnace. 

480.  The  tubes  which  support  the  brick  arches  are  threaded  at  each  end,  12  threads 
per  inch ;  the  ends  are  fitted  in  a  nut  shown  in  Fig.  717 ;  the  outside  of  the  nut  is 

tapered  and  threaded.  In  placing  one  of  the 
supporting  tubes  in  position,  the  nuts  are 
first  screwed  on  the  ends  of  the  tube,  and 
then  the  nuts  are  screwed  into  the  furnace 
sheets  until  a  steam-tight  joint  is  secured. 
These  nuts  are  made  either  of  wrought-iron 
or  brass. 


WOOTTEN  BOILER. 

481.  The  Wootten  boiler,  shown  in  Figs. 
718  and  719,  was  designed  by  John  E.  Woot- 
ten, General  Manager  of  the  Philadelphia  and 
Beading  R.  R.,  and  patented  by  him  July  1, 
1877.  This  boiler,  we  believe,  was  originally 
designed  for  the  purpose  of  burning  finely 
broken-up  coal,  generally  called  "  Buck- 
wheat," which  could  not  be  burnt  advan- 
tageously in  any  of  the  ordinary  locomo- 
tive boilers,  and  for  this  purpose  the  Wootten 
boiler  is  well  adapted.  Afterwards  the  ordi- 
nary anthracite  and  bituminous  coals  were 
also  burnt  in  this  boiler,  but  it  seems  to  us 
that  with  this  fuel  better  results  could  have  been  obtained  if  it  had  been  possible 
to  increase  the  aggregate  cross-sectional  area  of  the  tubes  so  as  to  obtain  a  ratio  of 
tube  area  and  grate  surface  nearer  equal  to  that  in  the  large  ordinary  boilers  as  given 
in  Table  64. 

The  distinctive  features  of  the  Wootten  boiler  are  a  comparatively  wide  and  shallow 
fire-box,  a  combustion  chamber  A,  and  a  brick  bridge  B  extending  across  the  fire-box 
end  of  the  combustion  chamber.  The  width  of  the  fire-box  may  be  extended  as  far  as 
the  width  of  the  roadway  will  permit.  The  fire-box  is  placed  above  the  frames  and 
extends  over  the  rear  driving  axle,  consequently  the  fire-box,  like  those  of  other  boilers 
with  the  furnace  placed  above  the  frames,  must  of  necessity  be  made  shallow,  so  as 
not  to  raise  the  boiler  too  high  above  the  track.  In  a  Wootten  boiler  in  which  finely 
broken-up  coal  is  burnt  a  brick  arch  is  of  greater  necessity  than  in  other  boilers ;  it 
serves  to  prevent  the  fine  coal  from  being  drawn  into  the  tubes  when  the  proper  thick- 
ness of  fire  is  carried.  Quite  a  number  of  these  boilers  are  now  in  use,  and  for  burn- 
ing fine  coal  a  better  boiler  can  probably  not  be  found. 


Fig.  717 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


479 


STRESS  ON  STAY  BOLTS. 

482.  Before  wo  leave  the  sub- 
ject of  boiler  construction,  it  may 
be  advantageous  to  examine  the 
stress  on  stay  bolts  and  oblique 
braces.  Fig.  720  shows  a  number 
of  stay  bolts  placed  in  horizontal 
and  vertical  rows;  the  distance 
between  centers  is  4£  inches, 


Fig.1 720 


and  they  are  supposed  to  sus- 
tain flat  surfaces  similar  to  those 
in  a  fire-box.  It  is  now  required 
to  find  the  stress  on  each  stay 
bolt,  the  steam  pressure  being  150 
pounds  per  square  inch.  Each 
stay  bolt  has  to  support  a  portion 
of  the  surface,  whose  length  and 
breadth  is  equal  to  the  distance 
between  the  centers  of  the  stay 
bolts ;  for  instance,  the  bolt  s  will 
have  to  support  that  portion  of  the 
surface  which  is  shaded,  and  the 
sides  of  this  surface  pass  through 
the  center  of  the  spaces  between 
the  bolt  s  and  those  nearest  to  it, 
as  SH  $2,  s3,  and  s4 ;  the  sides  of 
this  surface  must  also  be  parallel 
to  center  lines  e  g  and  e  f  drawn 
through  the  horizontal  and  ver- 
tical rows,  consequently  the  stay 
bolt  s  has  to  support  a  square  sur- 
face whose  sides  are  4J  incln's  lon^. 
The  area  of  this  surface  is  4.25  x 
4.25  =  18.0625  square  inches ;  and 


480  MODERN  LOCOMOTIVE   CONSTRUCTION. 

since  the  steam  pressure  is  150  pounds  per  square  inch,  the  pressure  on  this  surface,  and 
therefore  the  stress  on  the  stay  bolt  s,  will  be  18.0625  x  150  =  2709.375  pounds.  This 
pressure  has  to  be  resisted  by  the  smallest  cross-section  of  the  bolt,  which  is  at  the 
bottom  of  the  thread.  Now  suppose  that  these  stay  bolts  are  &  inch  diameter,  and  at  the 
bottom  of  the  thread  they  are,  say,  £  inch  diameter,  then  the  area  of  the  smallest  cross- 
section  will  be  equal  to  that  of  a  circle  £  inch  diameter,  which  is  equal  to  .44  square 
inch.  Here,  then,  we  see  that  on  .44  square  inch  of  metal  the  stress  is  2709.375  pounds- 
In  practice  it  is  always  necessary  to  know  how  much  that  will  be  equivalent  to  per 
square  inch,  so  that  a  comparison  can  be  made  with  the  limit  of  the  stress,  which  is 
always  given  per  square  inch.  The  equivalent  stress  per  square  inch  is  readily  found 
by  the  rule  of  proportion,  and  the  statement  takes  the  following  form  : 

.44  :  1  :  :  2709.375  :  stress  per  square  inch, 
from  which  we  get 

2709.375 
~~TT~      =  6157.6  pounds  per  square  inch. 

The  limit  of  safe  stress  per  square  inch  of  stay  bolts  should  not  exceed  6,000  pounds, 
hence  we  see  that  the  stress  is  somewhat  too  great,  and  therefore  the  diameter  of  the 
bolt  should  be  made  larger,  or  the  stay  bolts  should  be  placed  closer  to  each  other. 
From  the  foregoing  we  can  establish  a  rule  for  finding  the  stress  on  the  stay  bolt. 

EULE  106.  —  Multiply  the  distance  between  two  stay  bolts  on  the  horizontal  row  by 
the  distance  between  two  stay  bolts  on  the  vertical  row,  and  multiply  this  product  by 
the  steam  pressure  per  square  inch.  Divide  this  last  product  by  the  area  in  square 
inches  of  the  smallest  cross-section  of  the  bolt;  the  quotient  will  be  the  stress  per 
square  inch. 

483.  The  foi-egoing  also  indicates  the  method  for  finding  the  distances  between 
the  stay  bolts.  For  example  :  What  should  be  the  distance  between  the  stay  bolts  for 
a  steam  pressure  of  150  pounds  per  square  inch,  the  area  of  the  smallest  cross-section 
being  .44  square  inch,  and  the  limit  of  the  stress  6,000  pounds  per  square  inch? 
Now,  with  a  limit  of  stress  of  6,000  pounds  per  square  inch,  that  on  .44  square  inch  is 
found  by  the  following  proportion  : 

1  :  0.44  :  :  6000  :  x, 

from  which  we  get  .44  x  6000  =  2640  pounds  on  an  area  of  .44  square  inch.  The 
steam  pressure  is  150  pounds  per  square  inch,  consequently  the  area  of  the  sur- 

2640 
face  which  the  bolt  can  support  is  equal  to        r  =  17.6  square  inches.    If,  now,  the 


vertical  distance  between  the  stay  bolts  is  equal  to  their  horizontal  distance,  and 
if  these  rows  are  perpendicular  to  each  other,  then  the  form  of  the  surface  which  each 
stay  bolt  has  to  support  will  be  a  squai'e,  and  the  sides  of  this  surface  will  be  equal  to 
the  squai'e  root  of  17.6  =  V17.6  =  4.19  inches,  which  is  the  distance  between  the  stay 
bolts.  From  the  foregoing  we  can  establish  the  following  rule  for  finding  the  distance 
between  the  centers  of  the  stay  bolts  : 

EULE  107.  —  Multiply  the  limit  of  the  stress  per  square  inch  (6,000  pounds)  by  the 
area  in  square  inches  of  the  smallest  cross-section  of  the  bolt,  and  divide  this  product 


MOI)KR\    I.OCOMOTll'K 

by  the  steam  pressure  per  square  inch  ;  the  quotient  will  be  the  area  in  square  inches 
which  the  bolt  can  support.  If,  now,  the  horizontal  distances  between  the  stay  bolts 
are  equal  to  the  vertical  distances,  then  find  the  square  root  of  the  above  quotient  ; 
the  result  will  be  the  distance  in  inches  between  the  centers  of  the  stay  bolts. 

THICKNESS   OF   STAYED   SHEETS. 

484.  So  far  we  have  assumed  that  the  sheets  through  which  the  stay  bolts  pass 
are  sufficiently  thick  not  to  bulge  between  the  stay  bolts.  If  we  have  any  doubt  as  to 
this  thickness,  we  can  find  out  what  it  should  be  by  the  following  rule  : 

RULE  108.  —  When  the  horizontal  distances  between  the  stay  bolts  are  equal  to 
the  vertical  distances,  as  shown  in  Fig.  720,  multiply  the  square  of  the  distance 
between  the  stay  bolts  by  the  steam  pressure  per  square  inch,  and  divide  the  product 
by  8,000  ;  multiply  this  quotient  by  f  ;  the  square  root  of  this  last  product  will  be 
the  thickness  of  the  plate  whose  tenacity  is  64,000  pounds  per  square  inch.  This  rule 
is  expressed  by  symbols,  as  follows  : 

p 


in  which  t  denotes  the  thickness  of  plate  in  inches  ;  a,  the  distance  between  the  stay 
bolts  ;  p,  the  steam  pressure  per  square  inch  ;  and  /  the  working  pressure,  which 
will  depend  on  the  tenacity  of  the  plate.  If  the  tenacity  is  64,000  pounds  per  square 
inch,  which  is  about  correct  for  steel  sheets,  and  adopting  a  factor  of  safety  of  8,  then 
the  value  of  /will  be  8,000,  as  given  in  the  rule. 

EXAMPLE  144.  —  What  should  be  the  thickness  of  a  steel  plate  (tenacity  of  64,000 
pounds  per  square  inch)  subjected  to  a  pressure  of  150  pounds  per  square  inch,  stay 
bolts  4J  inches  from  center  to  center  I 

Substituting  for  the  symbols  in  the  formula  their  values,  we  have 


t.252  x  150       n    .     , 

-  =  .27  inch. 


Since  the  sheets  in  the  fire-box  are  never  less  than  J  inch  thick,  and  often  are 
re  inch  thick,  we  conclude  that  the  stay  bolts  are  spaced  sufficiently  close  in  this  case 
to  prevent  the  sheets  from  bulging. 

STRESS   IN  OBLIQUE  BRACES. 

485.  Sometimes  the  flat  surfaces,  such  as  the  boiler  heads,  are  stayed  by  braces 
similar  to  that  shown  in  Fig.  722 ;  the  end  g  of  this  brace  is  riveted  to  the  boiler  shell, 
and  the  other  end  is  connected  to  a  crow-foot  h.  If,  now,  we  wish  to  find  the  stress  per 
square  inch  on  the  brace  B,  we  must  first  find  the  pressure  which  the  crow-foot  h  has  to 
resist.  Let  Fig.  721  represent  the  position  of  the  crow-feet  on  the  sheet.  Now,  assume 
that  the  horizontal  distance  between  center  of  one  crow-foot  and  that  of  the  next  one  is 
9  inches,  and  the  vertical  distance  from  the  center  of  one  row  to  the  next  one  is  44 
inches;  each  crow-foot  will  have  to  support  a  surface,  as  indicated  by  the  shaded 
portion  abed,  9  inches  long  and  4£  inches  wide.  If,  now,  th<>  pivssmv  is  ]."><)  pounds 


482 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


f 


per  square  inch,  then  the  total  pressure  which  the  crow-foot  will  have  to  resist  is 
equal  to  9  x  4.5  x  150  =  6075  pounds.  This  pressure  or  force  acts  in  a  direction 
perpendicular  to  the  sheet  C  D.  Therefore,  through  any  point  e  on  the  center  line  of 
the  brace  B  draw  a  line  e  f  perpendicular  to  C  D,  and  make  the  length  of  this  line  to 

represent  the  pressure  of  6,075  pounds. 
This  can  be  done  as  follows :  Let  any  unit  of 
measurement  represent  a  certain  number 
of  pounds — for  instance,  let  1  inch  repre- 
sent 1,000  pounds ;  under  these  conditions 
the  line  c  b  must  be  6.075  inches  long. 
Through  the  end /draw  a  line/^r  perpendicular  to  F  G,  meeting  the  center  line  of  the 
brace  in  the  point  g,  thereby  completing  the  triangle  e  f  g.  Now,  the  stress  on  the  brace 
will  be  represented  by  the  side  e  g  of  the  triangle ;  if  this  side  measures  7  inches, 
then  the  stress  on  the  brace  will  be  7  x  1000  =  7000  pounds,  because  we  have 
adopted  a  scale  of  1,000  pounds  per  inch.  If  the  limit  of  the  stress  is  6,000  pounds  per 


Fig.  721. 


Fig.  722 


square  inch,  then  the  cross-sectional  area  of  the  brace  will  have  to  be 


7000 
6000 


=  1.166 


square  inches,  and  the  corresponding  diameter  will  be  1  £  inches  nearly ;  hence  if  the 
brace  is  made  of  round  iron,  it  will  have  to  be  lj  inches  diameter. 


BOILER   FASTENINGS   TO   FRAMES. 

486.  Fig.  723  shows  the  ordinary  way  of  fastening  the  frames  to  a  fire-box  which 
extends  to  a  short  distance  below  the  frames.  Four  clamps  or  pads  are  attached  to 
each  side  of  the  fire-box.  They  are  made  of  wrought-iron  plates  %  to  $  inch  in 


Fig.  723 


thickness.  The  two  upper  pads  A  A  are  planed  to  fit  around  the  three  sides  of  the 
frame,  then  heated  and  fitted  to  the  boiler.  Liners  are  placed  between  the  inner 
sides  of  the  frames  and  outer  sides  of  the  fire-box,  after  which  the  pads  are  bolted  to 
the  boiler  by  studs  £  inch  diameter.  The  two  lower  pads  B  B  are  lipped  over  the 
outside  of  the  lower  frame  brace  and  fastened  to  it  by  two  bolts  c  c  through  each  pad. 
These  bolts  are  driven  into  the  frame  brace,  but  work  in  slots  cut  in  the  pads ;  these 
slots  are  cut  large  enough  to  admit  a  thimble  over  each  bolt.  These  thimbles  rest  on 
top  of  frame  brace  and  project  a  little  over  the  faces  of  the  pads,  so  that  the  plates, 


MODERX  LOCOMOTIVE   CONSTRUCTION. 


483 


which  are  bolted  rigidly  to  the  top  of  thimbles,  will  not  touch  the  pads,  and  give 
the  latter  freedom  to  move  in  the  direction  of  the  frame  brace  as  the  boiler  expands. 
The  pads  are  of  course  fastened  to  the  fire-box  in  a  manner  similar  to  that  of  tne 
upper  ones. 

487.  Fig.  724  shows  the  method  of  fastening  the  frames  to  a  fire-box  placed  above 
the  axles  and  extending  to  a  short  distance  below  the  tops  of  frames.  The  pads  B  7? 
are  bent  to  an  L  form,  and  are  fastened  to  the  fire-box  like  those  described  above 


\ 


Fig.  724 

They  are  prevented  from  moving  laterally  by  the  clamps  D  D,  which  embrace  the 
lower  parts  of  the  pads  and  are  bolted  to  the  top  of  frame,  giving  the  pads  sufficient 
freedom  to  move  longitudinally  as  the  boiler  expands. 

488.  Fig.  725  shows  one  method  of  fastening  the  frames  to  u  fire-box  placed  above 
the  frames,  the  width  of  the  fire-box  being  equal  to  the  distance  from  outside  to 
outside  of  frames.  In  this  case  a  stud  is  forged  to  each  pad  E  E,  and  corresponding 
studs  are  attached  to  the  side  of  the  frame.  The  links  C  C  passed  over  these  studs 


Fig.  725 

support  the  rear  end  of  the  boiler.  This  manner  of  fastening  the  frames  to  fire-box 
is  considered  to  be  defective  and  insecure,  and  some  master-mechanics  believe  it  to  be 
the  cause  of  breaking  the  frames.  A  better  way  is  to  fasten  the  pads  to  the  ends  of 
the  fire-box  and  allow  them  to  rest  on  transverse  braces  bolted  to  the  top  of  frames. 
The  pads  must  be  attached  to  these  braces  in  such  a  manner  as  to  allow  the  boiler 
freedom  to  expand. 

489.  In  order  to  obtain  a  good  hold  for  the  studs  through  the  pads,  a  plate  about 
g  inch  thick,  and  about  as  wide  and  long  as  the  pad,  is  placed  on  the  inside  of  the 
fire-box  shell  for  each  pad.  These  plates  are  not  riveted  to  the  shell,  they  are  held  in 
position  by  the  stay  bolts  which  are  screwed  into  them. 


484 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


GKATE-BAES. 

490.  Grate-bars  may  be  divided  into  three  classes ;  one  class  embraces  those 
designed  for  burning  wood ;  the  second  class  embraces  those  for  burning  bituminous 
or  soft  coal ;  and  in  the  third  class  we  may  place  the  grate-bars  for  burning  anthracite 
or  hard  coal. 

Figs.  726,  727,  and  728  show  different  views  of  grate-bars  A  designed  for  burning 
wood.  They  are  generally  made  of  cast-iron,  cast  in  groups  of  two  or  three  bars,  and 
rest  on  wrought-iron  bearers  B  which  are  supported  by  the  straps  or  hangers  C;  this 
p.  736  type  of  hanger  is  only  used  in  furnaces 

which  do   not  have  a  fire-box  ring;   in 
furnaces  which  have  a  fire-box  ring,  the 

Kg.  728 

&'! 


Fig.  727 


bearers  are  made  of  cast-iron,  and  bolted  to  the  bottom  of  the  ring.  The  grate-bars 
are  generally  £  inch  thick  at  the  top  and  tapered  to  about  £  inch  at  the  bottom;  the 
space  between  the  bars  at  the  top  is  usually  §  or  f  inch  wide. 

Occasionally  we  find  the  grate-bars  made  of  wrought-iron  bolted  together  in 
groups  of  two,  three,  or  four  bars. 

491.  The  depth  e  at  the  center  of  the  grate-bar  should  have  a  certain  proportion 
to  its  length.     The  following  is  a  good  rule : 

RULE  109. — For  cast-iron  bars,  multiply  the  square  root  of  the  length  in  inches 
by  the  decimal  .6 ;  the  product  will  be  the  depth  in  inches  at  e.  Expressing  this  rule 
in  symbols,  we  have  .6  V  length  in  inches  =  depth  at  center.  For  wrought-iron  bars, 
we  have  .5  -/length  in  inches  =  depth  at  center.  The  depth  d  at  the  ends  of  the  bars 
should  never  be  less  than  one-half  of  the  depth  at  the  center.  The  thickness  at  the 
bottom  of  the  bars  should  be  about  £  of  the  thickness  at  the  top. 

492.  Figs.  728A  to  731  show  different  views  of  a  cast-iron  rocking  grate  designed 
for  burning  soft  coal.     Many  kinds  of  soft  coal  tend  to  clink,  and  thereby  prevent 
free  combustion.     Hence  for  burning  soft  coal  it  is  necessary  to  adopt  a  grate  to 
which  a  rocking  motion  can  be  imparted,  thereby  breaking  up  the  clinkers  and  keeping 
the  fire  clean.     There  are  many  types  of  rocking  grates  in  use ;  we  have  shown  only 
two  of  the  principal  types. 

The  grate-bar,  Figs.  728J.  and  729  (usually  called  a  finger  grate),  consists  of  a 
hollow  center-piece  A  with  projections  or  fingers  cast  to  each  side  of  it,  and  arranged 
so  as  to  make  the  fingers  of  one  grate  pass  between  those  of  the  next  grates,  leaving  a 
space  of  J  inch  or  f  inch  wide  between  the  fingers.  These  center-pieces  A  and  pivots 
B  are  cast  in  one  piece ;  the  pivots  rest  in  indentations  cast  in  the  side  or  bearing 
bars  C,  which  are  bolted  to  the  bottom  of  the  fire-box  ring.  A  finger  at  one  end  of 
each  grate-bar  projects  downwards  and  is  formed  to  act  as  an  arm,  as  shown  at  D ; 


MODES*'  LOCOMOTirE   CONSTRUCTION. 


485 


486  MODERN  LOCOMOTIVE   CONSTRUCTION. 

all  these  arms  are  attached  to  the  wrought-iron  bar  E.  The  rear  end  of  this  bar  is 
attached  to  the  wrought-iron  link  F,  and  this  in  turn  is  connected  to  the  bottom  of 
the  shaking  lever  H.  This  lever  extends  upwards  through  the  foot-plate  and  projects 
about  6  to  9  inches  above  it,  and  works  on  the  fulcrum  fastened  to  the  bracket  G. 

In  order  to  obtain  sufficient  power  for  rocking  the  grates,  the  upper  end  of  the 
shaking  lever  H  is  made  to  fit  in  the  socket  of  a  handle  (not  shown)  which  is  usually 
about  2  feet  or  2  feet  6  inches  long,  and  when  not  in  use  it  is  put  out  of  the  way.  It 
will  be  seen  that  with  this  arrangement  a  rocking  motion  is  imparted  simultaneously 
to  all  the  grate-bars.  After  the  fire  has  been  cleaned  the  grate-bars  are  prevented 
from  turning  by  inserting  a  pin  through  the  shaking  lever  and  a  wrought-iron  bracket 
fastened  for  this  purpose  to  the  top  of  the  foot-plate. 

A  drop-plate  J  is  placed  at  the  front  of  the  furnace  for  the  purpose  of  dumping 
the  fire.  This  drop-plate  works  on  pivots  similar  to  those  on  the  grate-bars.  The 
drop-plate  is  held  up  in  position  or  allowed  to  drop  by  the  wrought-iron  arms  K 
forged  to  a  shaft  which  extends  across  the  furnace,  and  works  in  bearings  fastened  to 
the  bottom  of  the  furnace  or  engine  frames.  This  shaft  is  operated  by  the  wrought- 
iron  lever  J",  which  is  generally  held  in  position  by  a  bracket  L  (Fig.  731).  The  lever 
J  is  placed  outside  of  the  engine  frame  on  the  fireman's  side.  With  the  type  of  lever 
here  shown,  the  fireman  has  to  step  off  the  engine  in  order  to  dump  the  fire.  Some- 
times a  long  handle  extending  upwards  through  the  running  board  is  attached  to  the 
end  of  the  lever  J,  so  that  the  drop-plate  can  be  raised  or  dropped  without  stepping  off 
the  engine.  Drop-plates  as  here  shown  are  used  in  soft-coal  burning  engines,  and 
seldom  in  hard-coal  burners. 

It  may  be  here  stated  that  the  cast-iron  bracket  G  also  supports  at  its  lower  end  a 
fulcrum  for  a  bell  crank  for  operating  the  front  damper  as  explained  in  Art.  495. 

Figs.  732  to  738  show  different  views  and  details  of  another  type  of  rocking  grate. 
The  grates  are  operated  in  a  manner  similar  to  that  of  operating  the  grates  shown  in 
Figs.  728A  and  729,  and  need  no  further  explanation,  excepting  to  state  that  the 
drop-plate  is  operated  by  the  handle  H2,  which  is  placed  next  to  the  shaking-lever 
handle  H  shown  in  Fig.  738. 

The  bar  which  connects  the  handle  H2  to  the  drop-plate  coincides  in  Fig.  732  with 
the  bar  which  connects  the  cast-iron  grates,  and  it  may  therefore  appear  that  the  drop- 
plate  and  grate-bars  are  all  connected  to  one  bar ;  but  an  examination  of  the  drawing 
will  show  that  this  is  not  the  case. 

WATEE  GRATES. 

493.  Fig.  739  shows  an  elevation ;  Fig.  740,  a  half  plan ;  and  Fig.  741,  an  end 
view  of  a  water  grate.  This  type  of  grate  is  almost  universally  used  for  burning  hard 
coal.  It  consists  of  wrought-iron  tubes  A  A,  1&  or  2  inches  outside  diameter,  made 
extra  heavy,  the  thickness  being  about  %  of  an  inch.  It  will  be  noticed  that  these 
tubes  are  placed  in  an  inclined  position,  lower  at  the  front  than  at  the  back  end; 
the  inclination  is  about  £  inch  to  the  foot ;  and  one  of  the  objects  in  placing  these 
grates  in  this  position  is  to  obtain  a  continuous  circulation  of  water  through  them,  so 
as  to  prevent  them  from  being  burnt. 


MODEKN  LOCOMOTIVE   CONSTBVCTION. 


487 


488 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


MODERN  LOCOMOTIVE   CONSTRUCTION.  489 

The  front  ends  of  these  tubes  are  threaded,  12  threads  per  inch,  and  screwed  into 
the  tube  sheet  or  front  sheet  of  the  furnace,  as  shown  in  Fig.  743.  A  small  portion  of 
the  other  ends  of  the  tubes  are  turned,  with  wrought-iron  ferrules  D  fitted  over  them, 
and  copper  ferrules  E  are  placed  on  the  outside  of  the  wrought-iron  ones,  as  shown  in 
Fig.  743.  These  ends  of  the  tubes  are  inserted  into  corresponding  holes  in  the  furnace- 
door  sheet ;  the  wrought-iron  and  copper  ferrules  are  then  driven  in  tightly,  and  the 
whole  expanded  on  the  water  space  side  of  the  furnace-door  sheet.  Sometimes  cast- 
iroii  ferrules  in  place  of  the  wrought-irou  ones  are  used ;  separate  views  of  a  cast-iron 
ferrule  are  shown  in  Fig.  744,  and  separate  views  of  the  copper  ferrules  are  shown  in 
Fig.  745. 

In  Fig.  744  it  will  be  noticed  that  the  end  of  the  cast-iron  ferrule  is  cut  open 
u  short  distance  on  one  side  only ;  care  must  be  taken  that  this  cut  does  not  extend 
beyond  the  copper  ferrule  when  placed  in  position ;  the  object  of  the  cut  is  to  make 
a  tight  joint  when  the  ferrule  is  driven  home,  as  it  cannot  be  expanded.  Wrought- 
iron  ferrules  are  left  solid. 

In  some  cases  nothing  but  a  tapered  copper  ferrule  is  used,  as  shown  in  Fig.  742, 
which  is  expanded  with  the  tube  on  the  water  space  side  of  the  sheet. 

The  brass  plugs  C  C  (Figs.  739,  740,  and  741)  are  for  the  purpose  of  closing  up 
the  holes  in  the  back  head ;  these  holes  are,  of  course,  required  for  passing  the  tubes 
into  the  furnace ;  they  also  afford  an  opportunity  for  expanding  the  ends  of  the  tubes. 

In  many  engines  the  water  tubes  are  arranged  as  indicated  by  the  brass  plugs 
C  C  in  Fig.  741 — that  is  to  say,  some  of  the  tubes  are  placed  above  the  others ;  but  this 
is  by  no  means  a  universal  practice,  frequently  we  find  all  of  them  placed  in  one  plane. 

494.  The  ferrules  B  B,  Figs.  740,  741,  are  usually  made  of  ordinary  2-inch  boiler 
tubes ;  they  form  the  openings  through  which  the  dead  bars  are  passed.  These  bars 
are  made  of  solid  wrought-irou  about  If  inches  diameter.  They  have  usually  an  eye 


forged  to  one  end,  which  projects  beyond  the  back  end  of  the  fire-box  and  provides 
means  for  pulling  the  dead  bars  out.  They  extend  to  within  1  inch  or  less  of  the  front 
sheet  of  the  furnace,  and  are  supported  at  the  front  end  by  the  cast-iron  bearer  A' 
(Fig.  739);  separate  vi<-\vs  of  this  liran-r  are  given  in  Figs.  746,  747,  and  74S. 

Another  bearer  /  similar  in  form  to  that  of  the  bearer  K  is  used  for  supporting 
the  center  of  the  dead  liars,  and  also  the  water  tubes.  Sometimes  the  front  bearer  K 
supports  only  the  ends  of  the  dead  bars;  in  such  cases  the  indentations  for  the  watrr 
grates  shown  in  Fig.  746  are  left  off;  but  the  central  bearing  7  always  supports  the 
tubes  as  well  as  the  dead  bars. 

Some  master-mechanics  consider  it  dangerous  pi'actice  to  put  the  plugs  C  C  in  the 


490  MODERN  LOCOMOTIVE   CONSTRUCTION. 

back  head,  because  should  one  of  these  plugs  blow  out  the  engineer  and  fireman  are 
liable  to  be  seriously  injured.  We  therefore  meet  with  engines  in  which  the  water 
tubes  have  been  passed  into  the  furnace  through  openings  in  the  front  end,  bringing 
the  plugs  C  C  in  the  front  sheet ;  hence,  in  this  case,  the  bottom  of  the  back  head  has 
no  other  large  holes  excepting  those  for  the  openings  through  which  the  dead  bars  are 
passed  into  the  furnace.  With  this  arrangement  the  water  tubes  are,  so  to  speak, 
turned  end  for  end,  the  threaded  ends  of  the  tubes  are  screwed  into  the  furnace-door 
sheet,  and  the  ferrules  on  the  tubes  are  driven  into  the  front  sheet  of  the  furnace. 
This  manner  of  placing  the  water  tubes  in  the  fire-box  is  not  free  from  objections,  as 
it  makes  the  cleaning  of  the  tubes  very  unhandy  and  difficult. 


CHAPTEK    XI. 


ASH-PANS.— SMOKE-STACKS— EXHAUST-PIPES. 


ASH-PANS. 

495.  Fig.  749  shows  an  elevation ;  Fig.  750,  part  of  a  plan ;  and  Fig.  751,  part  of 
an  end  view  of  the  simplest  kind  of  an  ash-pan.  This  type  is  used  for  boilers  whose 
fire-boxes  extend  downwards  between  the  driving  axles,  consequently  it  is  used  for 
eight-wheeled  passenger  engines,  Mogul,  and  ten-wheeled  engines. 

The  function  of  an  ash-pan  is  twofold,  namely,  to  concentrate  the  current  of  air 
as  it  flows  into  the  furnace,  and  afford  means  (but  it  does  not  always  do  so)  for  regu- 
lating a  supply  of  air  to  the  furnace. 

The  other  function  is  the  collecting  and  holding  a  quantity  of  ashes  without 
choking  the  draft  and  burning  the  grate-bars.  To  meet  these  requirements  the  depth 


—wl-^ 

Eig.  7.4Q 


Fig.  751 


Fig.  750 


Fig.  752 


of  this  ash-jian  is,  in  large  engines,  generally  12  inches,  and  in  small  engines,  10 
inches,  seldom  less  than  9  inches.  The  width  of  the  ash-pan  is  generally  equal,  or 
nearly  so,  to  the  outside  width  of  the  fire-box.  It  is  fastened  by  the  hangers  A  (two 
on  each  side)  to  the  lower  brace  B  of  the  engine  frame  as  shown  in  the  illustrations. 


\ 

, 

1 

j 

4 

1 

3 

i 

jut 

f 

. 
f 

ji 

| 

i 

: 

) 

fm.  FhjeShee 
^\  

'^i^r\jr~' 

g 


ifODERX  LOCOMOTIVE  CONSTRUCTION. 


493 


Sometimes  the  width  of  the  ash-pan  is  equal  to  the  width  of  the  furnace ;  in  cases  of 
this  kind,  an  angle  iron  is  riveted  to  each  side  of  the  ash-pan,  which  is  fastened  to  the 
bottom  of  the  fire-box  ring  by  studs  with  nuts,  or  studs  with  keys  which  wedge  the 
angle  iron  to  the  bottom  of  the  furnace  ring. 

The  ash-pan  shown  in  Figs.  749,  750  is  made  of  wrought-iron.  All  the  sheets  are 
made  of  the  same  thickness,  No.  7  or  No.  8  Birmingham  wire  gauge.  In  order  to 
stiffen  the  pan,  two  straps  C  C,  l£  x  ^  inch,  are  riveted  to  the  bottom  and  sides,  and 
half-round  bands  are  riveted  to  ends  of  the  ash-pan.  The  ash-pan  doors  or  dampers 
are  hinged  to  the  fire-box.  The  back  damper  is  operated  by  a  handle  extending 
through  the  foot-plate.  This  handle  is  connected  to  the  jaw  E,  which  is  riveted  to  the 
damper.  The  front  damper  is  connected  to  a  reach-rod,  one  end  of  which  works  on  the 
pin  F  (Fig.  752) ;  the  other  end  is  connected  to  the  bell  crank  suspended  from  the  lower 
pin  N  in  the  bracket  G,  Fig  728A ;  the  bell-crank  is  in  turn  connected  to  a  handle 
extending  through  the  foot-plate  alongside  of  the  back  damper  handle.  The  whole 
arrangement  is  shown  in  Fig.  607,  and  it  will  there  be  seen  that  both  dampers  are 
operated  from  the  foot-plate.  No  provision  for  dropping  the  ashes  is  made  in  this 
type  of  ash-pan ;  the  ashes  must  be  raked  out. 

496.  A  similar  form  of  ash-pan  is  shown  in  Figs.  753,  754,  755.  A  number  of 
pipes  are  placed  cross-ways  close  to  the  bottom  of  this  pan,  for  cleaning  it  out  by 


S.-K.SM.. 


n  n 

— ttpjaagSMKSTTjrtjy,- 


Door  weighted 
toSIbe. 


_J 


Fig.  758 


Fig.  760 


Fig.  761 


Fig.  759 


means  of  steam  jets.  The  arrangement  is  shown  so  plainly  that  further  explanation 
is  unnecessary.  This  device  is  patented. 

Figs.  758,  759,  760,  and  761  show  different  views  of  an  ash-pan  with  side  doors  for 
cleaning  it. 

497.  Figs.  762,  763,  and  764  show  different  views  of  a  wrought-iron  ash-pan  for  a 
consolidation  engine.  This  pan  is  made  in  three  sections,  and  their  forms  are  such  as 
to  clear  and  protect  the  driving  axles  from  ashes  falling  upon  them.  The  ends  of 
these  sections  often  overlap  each  other  about  3  or  4  inches,  the  lap  being  over  the 
axles.  Sometimes  the  ends  butt  against  each  other,  and  the  joints  are  covered  with 
plates  which  are  fastened  to  the  sections  by  set  screws.  Tt  is  not  necessary  to  place 
the  joints  of  the  sections  directly  over  the  center  of  axles  as  here  shown ;  indeed,  these 


L-. 


O 


J9 


494  MODERN  LOCOMOTIVE  CONSTRUCTION. 

joints  cannot  always  be  placed  there,  because  the  length 
of  the  middle  section  must  be  adjusted  so  as  to  permit 
it  to  be  placed  under  the  engine  with  the  driving  axles 
in  position,  and  therefore  the  middle  section  will  fre- 
quently not  reach  from  center  to  center  of  axles,  and 
of  course  the  end  sections  will  have  to  be  made  longer 
so  as  to  meet  it. 

Frequently,  ash-pans  of  this  class  have  no  dampers 
for  regulating  the  supply  of  air  to  the  furnace.  In 
the  ash-pan  here  shown  the  air  is  admitted  through 

i the  openings  A,  31  x  8 

inches,  cut  in  the  ends 
of  the  ash-pan  near  the 
top.  If  these  openings 
are  found  to  be  too  large 
for  the  average  work 
which  the  engine  has  to 
do,  they  are  partly  cov- 
ered by  the  movable 
plate  fastened  to  each 
end  of  the  ash-pan. 

Each  section  of  the 
ash-pan  is  provided  with 
a  drop-plate  B;  these 
plates  work  on  pivots 
placed  centrally  on  the 
!  ends  of  the  plates. 
b  The  pivots  rest  in 
bearings  in  the  cast-iron 
frame  riveted  to  the  bot- 
tom of  each  section.  The 
arms  C  of  the  drop-plates 
are  connected  to  a  reach- 
rod,  and  this  rod  is  con- 
nected to  suitable  levers 
and  handles,  so  that  all 
drop-plates  can  be  turned 
simultaneously  from  the 
foot-plate  and  allow  the 
ashes  to  drop.  Some- 
times the  drop-plates  are 
made  to  slide  instead  of 
turning  them  on  pivots. 
498.  Figs.  765,  766,  and  767  show  different  views  of  a  cast-iron  ash-pan  similar  in 
form  to  the  one  which  has  just  been  described. 


496 


MODERN  LOCOMOTIVE    CONSTRUCTION. 


The  thickness  of  the  metal  is  i  inch.  The  sides  of  each  section  have  flanges 
cast  to  them ;  the  flanges  are  turned  outwards  and  are  bolted  to  the  end  plates.  This 
arrangement  prevents  the  bolts  from  being  burnt.  Although  this  ash-pan  passes  over 
one  axle  only,  it  is  nevertheless  made  in  three  sections  for  the  convenience  of  handling 


SS-diarr— 


- 

Fig.  769 


'Fig..  771 


and  dropping  the  ashes.  The  drop-plates  in  this  case  are  made  to  slide,  and  are  worked 
simultaneously  from  the  foot-plate.  This  ash-pan  has  no  dampers  for  regulating  the 
supply  of  air. 

SMOKE-STACKS. 

499.  There  are  a  great  variety  of  smoke-stacks.  Their  design  depends  chiefly  on 
the  kind  of  fuel  used.  The  duty  of  a  locomotive  smoke-stack  is  not  only  to  carry  off 
the  products  of  combustion,  but  it  also  serves  as  a  passage-way  in  which  the  action  of 
the  exhaust  is  enabled  to  create  a  draft.  Some  smoke-stacks  are  also  designed  for 
spark  arresters,  and  this  requirement  is  the  cause  of  so  many  different  forms  of  stacks. 

Fig.  769  shows  an  ordinary  stack,  usually  called  the  diamond  stack.  It  is  used  on 
soft-coal  burning  engines  with  short  smoke-boxes.  A  cast-iron  ring  A  is  riveted  to 
the  bottom  of  the  cylindrical  shells,  and  this  ring  is  fastened  by  four  or  six  studs  to 
the  cast-iron  saddle  B  B,  which  is  in  turn  bolted  by  four  or  six  £  or  1  inch  bolts  to  the 
top  of  smoke-box.  The  outer  chipping  strip  c  is  fitted  closely  to  the  smoke-box,  and 


M(WKK\  i.oco.MorifK  coysTiircTJox.  497 

the  inner  strip  d  is  allowed  to  project  into  the  box  for  the  pui'poso  of  preventing  the 
condensed  steam  from  running  along  the  outside  of  the  box. 

The  cylindrical  part  D  of  the  stack  often  consists  of  two  shells,  leaving  an  annu- 
lar spare  about  jf  inch  wide  between  them.  Sometimes  four  1-inch  holes  are  drilled 
through  the  outer  shell  just  above  the  flange  A,  and  another  four  holes  are  drilled 
through  the  outer  shell  near  the  top,  for  the  purpose  of  creating  a  circulation  of  air 
through  the  annular  space.  This  arrangement  prevents  the  outer  shell  from  becoming 
overheated  and  blistering  the  paint.  The  outer  shell  also  serves  another  purpose :  it 
does  not  permit  direct  contact  between  the  outer  air  and  the  inner  shell,  thereby 
preventing  to  some  extent  condensation  of  the  exhaust  steam,  and  consequently 
tending  to  maintain  the  force  of  the  draft.  The  thickness  of  the  inner  shell  is  about 
No.  8  Birmingham  wire  gauge,  and  the  thickness  of  the  outer  shell  when  painted  is 
about  No.  10  B.  W.  G. ;  and  when  the  outer  shell  is  not  painted  it  is  made  of  Eussia 
iron. 

A  wrought-iron  ring  is  placed  at  the  top  between  the  two  shells  and  riveted  to 
them  and  also  to  the  lower  hood  E.  The  netting  is  placed  between  the  two  hoods 
E  and  F,  and  the  whole  bolted  together. 

A  cast-iron  cone  or  deflector  G  is  supported  by  three  cone  bolts  H,  which  are 
riveted  to  the  inner  shell.  The  threaded  parts  of  these  bolts  are  of  sufficient  length  to 
adjust  to  some  extent  the  position  of  the  cone  G.  The  purpose  of  this  cone  is  to 
arrest  the  sparks  or  unconsumed  fuel  and  prevent  injury,  as  much  as  possible,  to  the 
netting.  In  a  few  cases  the  whole  stack  has  been  made  of  cast-iron,  but  these  have  not 
given  satisfaction,  as  they  are  too  heavy  to  handle,  and  are  liable  to  break  off.  The 
cylindrical  part  of  the  stack  does  not  always  consist  of  two  shells,  as  will  be  seen  in 
Fig.  770. 

500.  Figs.  770  and  771  show  an  elevation  and  part  of  the  plan  of  another  diamond 
stack.     The  principal  difference  between  this  stack  and  the  one  previously  described  is 
that  the  former  has  in  addition  to  the  cast-iron  cone  A  another  spark  arrester  B,  whose 
form  is  that  of  a  cone-shaped  spiral ;  it  is  made  of  round  bar  iron,  and  is  supported  by 
wrought-iron  brackets.     This  spiral  cone  cannot  be  adjusted  after  it  has  been  placed 
in  position. 

The  lower  and  upper  hoods  with  the  netting  between  them  are  clamped  together 
by  an  angle  iron  partly  closed  to  shape.  The  ends  of  this  angle  iron  are  turned  out- 
wards and  drawn  together  by  a  link  or  stirrup,  as  shown  at  6'  in  Fig.  771. 

501.  Fig.  772  shows  another  smoke-stack  for  a  soft-coal  burning  engine  with  short 
smoke-box. 

Fig.  773  shows  the  foregoing  stack  changed  to  suit  a  wood-burning  engine  with 
short  smoke-box.  The  whole  change  consists  in  the  form  of  the  upper  hood. 

502.  The  common  type  of  stack  for  wood-burning  engines  is  shown  in  Fig.  774. 
Tin-  annular  space  A  between  the  two  cylindrical  shells  is  designed  for  a  receptacle  for 
sparks,  and  is  therefore  made  much  wider  than  the  annular  space  shown  in  Fig.  7(59. 

The  outer  shell  is  provided  at  the  bottom  with  a  hand-hole  casting  and  cover  (not 
shown  here)  for  taking  the  cinders  and  ashes  out  of  the  annular  spare.  The  ordinary 
cone  or  deflector  is  also  used.  The  upper  hood  is  made  of  netting  (',  stiffened  by  the 
straps  d.  No  cast-iron  flange  at  the  bottom  of  the  cylindrical  shells  is  used ;  the  latter 


498 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


are  bolted  directly  to  the  saddle  G,  the  outer  shell  being  generally  bolted  to  the  saddle 
by  three  bolts  e,  and  the  inner  shell  bolted  to  the  same  by  three  bolts/ 

503.  Fig.  775  shows  the  celebrated  Eadley  &  Hunter  stack,  which  in  former  times 
was  the  favorite  for  wood-burning  engines.     It  has  no  netting,  but  in  place  of  it  a 


^=^^_l 


Fig.  774 


Fig.J775 


MODEBX  LOCOMOTIVE   CONSTRUCTION.  499 

wrought-iron  perforated  liner  A  is  used ;  the  holes  are  of  a  rectangular  form,  and  are 
here  only  indicated  in  section  at  b  b.  The  cast-iron  cone  C  has  supports  cast  to  it 
which  rest  on  the  wrought-iron  plate  d  d  riveted  to  the  inner  shell,  and  steadied  by  the 
wrought-iron  bracket  e  c. 

504.  Figs.  776  and  777  show  an  elevation  and  part  plan  of  another  wood-burning 
stack.     Its  main  feature  is  the  damper  on  the  top  of  stack ;  this  damper  has  connec- 
tions leading  to  the  cab,  so  that  it  can  be  opened  or  closed  from  there.    This  stack  is 
chiefly  used  for  plantation  districts,  and  when  running  through  or  near  these  districts 
the  damper  A  is  closed  so  as  to  avoid  setting  fire  to  the  crops. 

505.  Fig.  778  shows  a  straight  stack,  and  Figs.  779  and  780  show  a  section  and 
plan  of  its  saddle  and  flange.     This  stack  is  used  either  for  soft  or  hard  coal,  but  since 
it  has  no  netting  for  arresting  the  sparks  it  is  suitable  only  for  extension  fronts  or 
such  smoke-boxes  as  contain  a  netting.     The  cast-iron  top  consists  of  three  pieces, 
namely,  the  ring  (7,  the  flare  B,  and  the  cap  A.     The  ring  C  rests  on  the  top  of  the 
outer  shell,  the  flare  B  rests  on  this  ring,  and  the  cap  A  is  placed  on  the  top  of  the 
flare ;  these  are  not  riveted  nor  bolted  together,  but  they  are  firmly  held  together  by 
expanding  the  top  of  the  inner  shell  over  the  flange  on  the  cap  A. 

506.  Fig.  781  shows  a  stack  whose  outer  form  is  similar  to  that  of  the  one  we  have 
just  described,  but  the  form  of  the  inner  shell  of  this  stack  differs  from  the  foregoing. 
In  Fig.  781  the  diameter  of  the  top  of  the  inner  shell  is  much  larger  than  the  diameter 
at  a  short  distance  above  the  base ;  we  may  therefore  class  it  with  the  so-called  tapered 
stacks. 

Mr.  William  S.  Hudson  introduced  stacks  similar  to  that  shown  in  Fig.  781  on 
engines  built  by  the  Rogers  Locomotive  Works,  Paterson,  N.  J.,  some  twenty-five  years 
ago,  hence  the  tapered  stack  shown  in  the  next  figure  cannot  be  considered  to  be  a 
modern  improvement,  as  it  is  frequently  said  to  be,  although  the  tapei'ed  stack  is 
now  used  to  a  greater  extent  than  formerly. 

507.  Fig.  782  shows  a  tapered  stack.    The  saddle  is  made  of  cast-iron ;  the  stack 
is  usually  made  of  wrought-iron  about  J  inch  in  thickness.     Sometimes  the  stack  is 
made  of  cast-iron  about  fa  inch  thick.     Its  simplicity,  and  consequently  its  cheapness, 
recommends  its  adoption,  and  it  has  become  of  late  the  favorite  stack  for  hai'd-coal 
burning  engines ;  it  is  also  frequently  used  for  soft-coal  burners.     It  has  no  netting, 
and  it  is  therefore  suitable  only  for  extension  fronts,  although  on  one  of  our  prominent 
roads  it  is  used  for  short  smoke-boxes  which  have  no  netting;  but  such  practice  is 
not  to  be  recommended,  as  there  is  nothing  to  prevent  red-hot  cinders  or  unconsumed 
fuel  from  being  thrown  out  of  the  stack. 

Making  the  stack  larger  in  diameter  at  the  top  than  at  the  bottom  makes  it 
conform  more  to  the  path  of  the  gases  than  that  of  a  straight  stack,  consequently 
friction  will  be  reduced;  enlarging  the  upper  end  also  reduces  the  velocity  of 
the  products  of  combustion  in  their  ascent,  consequently  the  height  to  which 
they  are  thrown  into  the  air  will  also  be  reduced — all  of  which  tends  to  lessen  the 
work  of  the  exhaust  steam,  and  enables  it  to  expend  its  force  in  its  legitimate  duty, 
that  of  creating  a  draft.  The  amount  of  taper  will  depend  on  the  kind  of  exhaust 
nozzle  used;  generally  the  increase  of  diameter  is  about  1  inch  or  a  little  more  per 
foot  in  height. 


500 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


---  so'-t-11-  —  •  -- 

"  J"7ih.lf,.,..j;, 


Pig.  782 


MODEIiX  LOCOMOTIVE   COXSTRUCTIOX. 


501 


If  the  diameter  at  the  bottom  of  the  stack  is  too  great,  soot  is  liable  to  collect  on 
the  inside  near  the  bottom  of  the  stack ;  on  the  other  hand,  if  a  correct  taper  has  been 
given,  the  exhaust  steam  will  keep  the  stack  clean  throughout  its  length. 

508.  Figs.  783  and  784  show  a  tapered  stack  with  a  variable  exhaust  damper 
applied.  This  damper  is  a  recent  invention  of  Mr.  H.  A.  Luttgens,  who  for  many 

years  has  been  the  chief  draftsman  at  the  Eogers 
Locomotive  Works,  Paterson,  N.  J.  The  damper  is 
very  simple  in  construction,  and  can  be  easily  ap- 
plied to  any  stack  by  removing  the  base  or  saddle 
and  putting  the  damper  in  its  place.  It  consists 
of  a  cast-iron  register-plate  B  resting  upon  a  turned 


Fig.  783 


Fig.  784 


surface  covered  with  an  improved  Chinese  paint ;  this  paint  is  heat  proof,  and  enters 
into  the  pores  of  the  cast-iron,  forming  a  hard  and  durable  protection  to  the  iron. 

The  register-plate  is  provided  with  a  number  of  openings  which  correspond  to  the 
apertures  in  the  base.  It  is  connected  by  two  links  to  the  lever  working  on  the 
fulcrum  F,  and  this  lever  is  worked  by  the  rod  Z),  which  extends  into  the  cab,  and  is 
so  arranged  that  either  the  engineer  or  fireman  can  open  or  close  the  apertures  in 
the  base.  In  the  plan  of  the  stack  these  apertures  show  partly  open,  and  in  this 
position  air  is  admitted  into  the  bottom  of  the  stack,  as  indicated  by  the  arrows  at  B 
in  the  elevation. 

With  this  device  the  engineer  is  enabled  to  regulate  the  draft  as  required,  by 
admitting  more  or  less  air,  or  shutting  it  off  at  will.  The  advantage  of  regulating  the 
draft  in  this  manner  will  readily  be  perceived ;  it  does  away  with  the  necessity  of 
opening  the  fire-door  and  admitting  a  current  of  cold  air  into  the  furnace,  which  often 
results  in  injury  to  the  furnace  sheets  and  tube  joints.  Evidence  also  seems  to 
indicate  that  with  this  damper  a  considerable  saving  of  fuel  can  be  secured. 

On  the  switching  engines  running  on  the  Long  Island  Railroad,  the  dampers  are 
closed  at  first  upon  a  new  fire,  but  during  the  remainder  of  the  day  each  engine  is  run 
with  an  open  damper;  the  result  claimed  for  this  is  a  noiseless  ami  smokeless  engine, 
without  waste  of  steam  at  the  safety-valve,  and  using  from  3,000  to  4,000  pounds  of 
inferior  coal,  where  3  tons  would  be  about  the  usual  allowance.  These  switching 


502  MODERN  LOCOMOTIVE   CONSTRUCTION. 

engines  have  18  x  24  inch  cylinders,  and  are  on  the  move  continuously.  In  other 
cases  it  is  claimed  that  a  saving  of  fuel  from  15  to  30  per  cent,  has  been  secured ; 
although  we  have  no  reason  for  doubting  these  statements,  it  seems  hardly  possible 
that  so  favorable  results  can  be  obtained  in  every  case.  Many  of  these  dampers  have 
been  put  in  practical  use  and  seem  to  establish  an  excellent  reputation.  Whatever 
merits  may  be  attributed  to  them,  it  is  evident  that  they  correct  some  defects  in 
the  steaming  of  a  locomotive.  In  the  southern  States  there  are  quite  a  number  of 
engines  running  with  these  dampers,  and  it  is  claimed  that  they  have  proved  them- 
selves valuable  on  account  of  their  spark-arresting  quality,  and  have  prevented  fires 
in  the  cotton  districts.  Undoubtedly,  as  the  adoption  of  these  dampers  extends  their 
advantages  will  be  more  appreciated. 

509.  In  extension  fronts  where  long  exhaust  nozzles  reach  nearly  to  the  top  of  the 
smoke-box,  some  attention  must  be  given  to  the  inner  shape  of  smoke-stack  saddle, 
because  the  free  access  to  the  stack  of  the  products  of  combustion  will  be  somewhat 
affected  by  this  shape.     If  the  entrance  of  the  saddle  into  the  smoke-box  is  made 
square,  as  shown  at  A  A  in  Fig.  772,  the  distance  between  the  top  of  the  exhaust 
nozzle  and  the  top  of  the  smoke-box  should  be  equal  to  about  the  smallest  inner 
diameter  of  the  stack.    When  the  entrance  of  the  saddle  into  the  smoke-box  is  of  a 
bell  form,  as  shown  at  B,  Fig.  782,  then  the  distance  from  the  top  of  the  exhaust 
nozzle  and  the  top  of  smoke-box  can  be  made  a  little  less.     In  every  case  care  must 
be  taken  not  to  impede  the  access  of  the  gases  to  the  stack,  and  the  steam  must  be 
discharged  through  the  gases  and  not  above  them. 

510.  Locomotive  stacks  cannot  be  made  long  enough  to  create  a  natural  draft, 
consequently  other  means  for  obtaining  the  required  draft  must  be  provided.     This  is 
accomplished  by  the  action  of  the  exhaust  steam,  which  may  be  described  briefly  as 
follows :  As  the  exhaust  steam  passes  through  the  stack  it  acts  like  a  piston  in  a 
cylinder,  taking  with  it  a  portion  of  the  gases  in  the  smoke-box.    The  blasts,  which 
occur  in  quick  succession,  create  a  partial  vacuum  in  the  smoke-box  and  furnace, 
causing  the  external  pressure  of  the  atmosphere  to  force  air  through  the  openings  in 
the  grate  and  through  the  fuel,  from  whence  the  air  and  products  of  combustion  are 
drawn  through  the  tubes  and  finally  ejected  through  the  smoke-stack  into  the  outer 
air.     From  the  foregoing  it  will  be  seen  that  the  stack  alone,  or  the  blast  without  the 
stack,  cannot  create  the  required  draft,  but  the  two  must  act  in  conjunction. 

DIAMETERS   OF  STACKS. 

511.  It  may  therefore  be  reasonably  assumed  that  the   stacks  should  have  a 
definite  size,  but  on  this  matter  engineers  do  not  seem  to  agree.     Our  experience 
seems  to  indicate,  and  is  partly  confirmed  by  practice,  that  the  internal  area  of  the 
smallest  cross-section  of  the  stack  should  be  -^  of  the  area  of  the  grate  surface. 

The  first  column  in  the  following  table  gives  the  sizes  of  cylinders ;  the  second 
column  gives  the  grate  area  in  square  inches  for  soft  coal  as  found  by  Eule  103,  and 
previously  tabulated  in  Tables  59  and  60 ;  the  third  column  gives  the  smallest  cross- 
sectional  area  of  the  stack  equal  to  -fa  of  the  grate  area ;  and  the  fourth  column  gives 
the  corresponding  diameter  in  the  nearest  ^  inch. 


MODERX  LOCOMOTIVE   CONSTRUCTION. 


503 


TABLE   73. 

CALCULATED   DIAMETERS  OP   STACKS. 


Size  of  Cylinders. 

Grate  Area. 

Cross-sectional  Area  at  the 
Smallest  Part  of  the  Stack. 

Diametera  of  Stacks. 

Column  1. 

Column  a. 

Column  3. 

Column  4. 

10  x  20  inches. 

950.40  sq.  in. 

55.90  sq.  in. 

8i  inches. 

11  x  22 

1267.20 

1 

74.54 

9J 

12  x  22 

HJ.l.eO 

1 

83.85 

10* 

13  x  22 

1627.20 

4 

95.71 

11 

14  x  24 

1843.20 

1 

108.42 

Hi 

15  x  24 

2116.80 

1 

124.51 

12| 

16  x  24 

•J-_'7f).20 

t 

133.83 

13 

17  x  24 

2491.20 

4 

146.54 

m 

18  x  24 

3750.44 

1 

161.77 

14| 

20  x  24 

4:120.00       " 

254.11 

18 

22  x  24 

5011.20      " 

294.77 

19f 

The  foregoing  diameters  of  stacks  have  been  determined  by  the  grate  area  of  soft- 
coal  burning  engines.  For  hard-coal  burning  engines  the  grate  area  is  larger,  and 
therefore  it  would  seem  that  the  diameters  of  stacks  should  also  be  greater  for  this 
class  of  engines ;  but  since  in  hard-coal  burning  engines  lighter  fires  are  carried,  the 
diameters  of  stacks  given  in  the  foregoing  table  need  not  be  changed. 

The  following  table  gives  the  diameters  of  stacks  in  actual  service.  A  comparison 
shows  that  the  calculated  diameters  agree  very  closely  with  the  diameters  of  stacks  in 
actual  use. 

TABLE  74. 

DIAMETERS  OF   STACKS  IN   ACTUAL  SERVICE. 


— 

Diameters  of  Cylinders. 

Stroke. 

Smallest  Inside  Diametera  of  Stacks. 

Column  1. 

Column  S. 

Column  :'.. 

9  inches. 

16  inches. 

7i  inches  to  8f  inches. 

10 

18      " 

8i 

11 

18       ' 

9± 

12 

20       ' 

B| 

to  10 

13 

20       ' 

11 

14 

20       ' 

11 

to  llf 

15 

22       ' 

m 

to  13                ami  14  inches. 

16 

24       ' 

13$ 

to  14                  "    14|     " 

17 

24       ' 

14± 

to  15                  "    15i     " 

18 

24      " 

14 

to  15                 "    16i     " 

19 

24      " 

16 

to  17 

20 

24       " 

17 

to  18 

21 

24      " 

18 

22 

24      " 

19 

512.  It  frequently  occurs  that  the  length  of  the  stack  is  limited  by  the  distance 
between  the  track  and  the  bridges  under  which  the  engine  has  to  pass,  consequently 
the  height  of  the  engine  measured  from  track  to  the  top  of  smoke-stack  is  limited  to 
14  feet  9  inches,  or  15  feet,  seldom  exceeding  the  latter.  But  in  cases  where  the 
height  of  the  stack  is  not  limited  by  the  bridges,  the  question  arises:  What  shall 
be  its  length?  Since  the  stack  cannot  be  made  high  enough  to  create  a  natural 


504 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


draft,  we  must  assign  to  it  a  length  suitable  for  the  action  of  the  exhaust  steam. 
Mr.  D.  K.  Clark,  in  his  treatise  on  Railway  Machinery,  recommends  a  length 
equal  to  four  diameters  of  the  stack.  This  proportion  we  believe  to  be  a  good  one, 
although  some  builders  adopt  a  greater  length  when  not  limited  by  the  conditions 
previously  given.  Hence,  the  length  of  a  tapered  stack,  shown  in  Fig.  782,  should  be 
equal  to  four  times  its  smallest  inside  diameter,  and  the  length  of  the  cylindrical  part 
of  the  diamond  stacks,  shown  in  Fig.  769,  should  be  equal  to  four  times  the  inner 
diameter  of  the  inner  shell,  and  the  height  of  the  inner  pipe  in  wood-burning  stacks 
should  also  be  equal  to  four  times  its  inner  diameter. 


EXHAUST   PIPES   AND   NOZZLES. 

513.  Figs.  785,  786,  787  show  different  views  of  a  double  exhaust  pipe,  and  Figs. 
788,  789  show  its  nozzle. 

This  pipe  is  comparatively  short ;  it  is  used  in  connection  with  stacks  which  have 
deflectors  or  cones,  and  netting  for  arresting  the  sparks  and  cinders ;  and  it  may  be 


said  that  this  pipe  is  generally  used  in  short  smoke-boxes  with  petticoat  pipes,  through 
which  the  exhaust  steam  passes  after  it  leaves  the  nozzle. 

The  exhaust  pipe  is  bolted  to  the  cylinder  saddle  by  studs  passing  through  the 
lugs  b  b,  and  the  nozzle  is  fastened  to  the  top  of  the  pipe  by  the  studs  a  a,  which  pass 
through  the  lugs  c  c  of  the  nozzle.  The  object  of  making  the  nozzle  separate  from  the 
pipe  is  to  provide  ready  means  for  replacing  it  by  a  larger  or  smaller  one,  so  that  the 
force  of  the  blast  can  be  decreased  or  increased  as  occasion  may  require.  Usually 
three  nozzles  of  different  sizes  are  sent  with  each  engine. 

The  openings  d  d,  Fig.  785,  at  the  bottom  of  the  pipe  coincide,  of  course,  with  the 
openings  in  the  cylinder  saddle,  and  the  openings  e  e  at  the  top  of  the  pipe  are  placed 
as  close  to  each  other  as  it  is  practical  to  do. 

The  openings  f  f,  Fig.  788,  at  the  top  of  the  nozzle  are  bored  out  cylindrically 
through  a  distance  of  about  1  inch,  so  as  to  prevent  as  much  as  possible  the  spreading 
of  the  exhaust  steam.  The  axes  of  these  openings  //  lie  in  a  transverse  plane  perpen- 


MODE11X  LOCOMOTIVE  CONSTRUCTION. 


505 


dicular  to  the  axis  of  the  boiler,  but  the  planes  parallel  to  the  axis  of  the  boiler  passed 
through  these  axes  of  the  openings  //  slightly  incline  towards  each  other  at  the  top, 
so  as  to  direct  the  blast  towards  the  bottom  center  of  the  cone  in  the  stack.  The 
passages  through  the  exhaust  pipe  gradually  decrease  in  size,  and  this  is  done  for  the 
purpose  of  increasing  the  velocity  of  the  exhaust  steam  as  it  passes  out  of  the  nozzle, 
thereby  giving  the  blast  a  sufficient  force  to  create  the  required  draft. 

514.  Figs.  790,  791,  792  show  another  short  exhaust  pipe ;  the  form  of  the  bottom 
of  this  pipe  is  similar  to  that  of  the  one  which  has  just  been  described,  but  the  top 
of  the  pipe  is  designed  for  one  orifice  only,  through  which  the  steam  from  both 


EXHAUST     NOZZLE. 


cylinders  is  discharged ;  this  pipe  is  therefore  called  a  single  exhaust  pipe.  A  thimble 
B  is  inserted  and  fastened  by  a  set  screw  in  the  top  of  the  pipe ;  the  object  of  using  a 
thimble  is  to  provide  means  for  changing,  in  an  easy  manner,  the  size  of  the  orifice  by 
using  a  larger  or  smaller  thimble.  Usually  three  thimbles  of  different  sizes  are  sent 
with  each  engine. 

The  upper  part  of  the  thimble  is  bored  out  cylindrically  and  must  point  direct  to 
the  center  of  the  cone  in  the  stack.     It  will  be  well  to  notice  the  outside  form  of  the 


506 


MODERN  LOCOMOTIVE   CONSTRUVTIOX. 


Fig.  79G 


thimble ;  it  is  tapered,  smaller  at  the  top  than  at  the  bottom,  and  the  upper  edge  of  the 
pipe  is  chamfered  off;  this  is  done  to  facilitate  the  approach  of  the  gases  to  the 
exhaust  steam.  For  similar  reasons  the  outside  of  the  double  nozzle  in  Figs.  788,  789 
should  have  been  tapered  more  than  as  is  there  shown. 

The  bridge  which  divides  the  pipe  at  the  bottom  into  two  passages  prevents,  or 
should  prevent,  a  flow  of  exhaust  steam  from  one  cylinder  into  the  other  one,  and 
consequently  the  bridge  in  these  short  pipes  must  be  made  as  high  as  possible. 

515.  It  is  important  that  the  passages  through  the  double  exhaust  pipe  and 
nozzle,  as  well  as  the  passages  through  the  single  exhaust  pipe  and  thimble,  should  be 
smooth,  continuous,  and  with  easy  curvature ;  all  sudden  bends  and  corners  should  be 
avoided,  and  the  cross-sections  of  each  passage  should  be  similar  throughout.  By 
attending  to  these  particulars  the  back  pressure  in  the  cylinders  will  be  reduced. 
Either  the  double  exhaust  pipe  or  the  single  one  can  be  used  in  one  and  the  same 
engine,  but  which  one  of  these  two  is  best  to  adopt  is  difficult  to  decide,  and  engineers 
do  not  agree  on  this  subject.  The  advocates  of  the  double  exhaust  pipe  claim  that  by 
its  use  the  exhaust  from  one  cylinder  cannot  possibly  interfere  with  the  exhaust  of 

the  other  one,  and  therefore  the  back  pressure  in 
either  cylinder  cannot  be  increased  by  the  exhaust 
steam  from  one  cylinder  flowing  back  into  the  other, 
as  they  say  is  liable  to  occur  with  the  single  exhaust 
pipe.  On  the  other  hand,  the  advo- 
s  thimbles  to  bore  2x#izx  cates  of  the  single  exhaust  pipe  claim, 
and  justly  so,  that  the  blast  should 
be  concentric  with  the  stack,  and  this 
can  only  be  obtained  with  a  single 
pipe.  In  order  to  prove  that,  with 
the  use  of  the  single  pipe,  the  exhaust 
steam  from  one  cylinder  cannot  flow 
back  into  the  other,  indicator  cards 
have  been  taken  with  each  kind  of 
pipe,  and  these  cards  do  not  show 
any  appreciable  difference  in  the  back 
pressure.  Our  choice  is  the  single  ex- 
haust pipe,  because,  we  believe,  if  it 
is  correctly  designed  it  will  give  good 
results,  and  it  is  easier  to  make  and  to 
keep  in  repair. 

516.  With  these  short  single  ex- 
haust pipes  a  petticoat  pipe  must  also  be  used ;  but  when 
a  diaphragm  plate  is  employed,  as  shown  in  Fig.  714,  the 
petticoat  pipe,  in  the  majority  of  cases,  is  dispensed  with, 
or  in  the  few  cases  where  it  is  used  it  is  very  short ;  conse- 
quently the  exhaust  pipe  used  in  connection  with  a  dia- 
phragm plate  is  made  much  longer  than  that  shown  in  Figs.  790,  791.  Since  diaphragm 
plates  are  employed  in  nearly  all  extension  fronts,  it  may  be  said  that  the  long  single 
exhaust  nozzle  belongs  to  the  extension  fronts. 


BX- 

Fig.  794 


MODESX  LOCOMOTIVE  CONSTRUCTION.  507 

517.  Figs.  793,  794,  795,  796  show  different  views  of  a  single  long  exhaust 
pipe ;  its  general  construction  does  not  differ  from  that  shown  in  Figs.  790  and 
791,  and  therefore  the  remarks  relating  to  the  latter  apply  equally  well  to  the  long 
pipe. 

These  pipes  should  be  made  as  long  as  possible,  leaving  only  a  sufficient  space 
between  the  top  of  the  thimble  and  the  bottom  of  the  stack  for  the  free  access  of  the 
gases  to  the  stack.  In  Art.  509  we  have  seen  that  when  the  stack  has  a  square 
entrance  to  the  smoke-box  the  distance  from  the  top  of  the  thimble  to  the  bottom  of 
the  stack  should  be  equal  to  the  diameter  of  the  stack ;  and  when  the  stack  has  a  bell- 
shaped  entrance,  the  exhaust  pipe  can  be  made  somewhat  longer.  These  conditions 
enable  us  to  determine  the  length  of  the  exhaust  pipe.  The  bridges  in  the  long 
exhaust  pipes  are  of  course  much  higher  than  those  in  shorter  ones,  and  therefore 
the  danger  of  the  exhaust  from  one  cylinder  flowing  back  into  the  other  is  mud: 
reduced,  and  single  pipes  can  be  used  with  confidence  in  extension  fronts.  Some 
master-mechanics  make  the  cross-section  of  the  body  of  the  exhaust  pipe  compara- 
tively small,  so  as  to  get  rid  of  the  exhaust  steam  as  quickly  as  possible,  thereby 
creating  a  strong  blast  writh  decisive  intervals.  Others  prefer  to  make  the  body  of 
the  pipe  comparatively  large,  and  contracting  it  near  the  top,  thereby  obtaining  a 
more  constant  and  milder  draft.  Either  plan  can  be  advantageously  employed ;  for 
shallow  fires  we  should  recommend  to  make  the  cross-section  of  the  body  of  the 
exhaust  pipe  comparatively  large,  so  as  to  obtain  a  mild  and  constant  draft  which  will 
not  tear  up  the  fire.  On  the  other  hand,  for  a  deep  fire  we  should  recommend  to 
reduce  the  cross-sectional  area  of  the  body  of  the  pipe,  so  as  to  obtain  a  blast  suffi- 
ciently strong  to  draw  the  air  through  the  increased  depth  of  fire. 


AREA   OF   ORIFICES   IN   EXHAUST   NOZZLES. 

518.  The  area  of  the  orifices  in  the  exhaust  nozzles  or  thimbles  will  depend  on  the 
quality  and  quantity  of  the  coal  to  be  burnt,  size  of  cylinder,  and  the  condition  of  the 
outer  atmosphere.  It  is  therefore  impossible  to  give  rules  for  computing  the  exact 
diameter  of  the  orifices.  All  that  can  be  done  is  to  give  a  rule  by  which  an  approxi- 
mate diameter  can  be  found,  which  may  then  be  used  as  a  basis  from  which  the 
exhaust  pipe  can  be  designed.  The  exact  diameter  of  orifice  can  only  be  found  by 
actual  trials,  and  such  sizes  should  be  adopted  as  will  give  the  greatest  economy  in 
fuel.  We  have  seen  that  among  the  things  which  affect  the  size  of  exhaust  nozzles  arc 
the  quantity  of  coal  to  be  burnt  and  the  size  of  cylinders,  and  these  two  things  affect 
its  size  more  than  anything  else.  The  quantity  of  coal  to  be  burnt  will  of  course 
depend  to  a  great  extent  on  the  grate  area,  and  since  this  area  is  made  proportional  to 
the  tractive  force  which  takes  in  account  the  size  of  cylinders,  we  may  at  once  deter- 
mine the  size  of  exhaust  nozzle  from  the  grate  area. 

Our  experience  leads  us  to  believe  that  the  area  of  each  orifice  in  double  exhaust 
nozzles  should  be  voo  Pai"t  of  the  grate  area.  Adopting  this  proportion,  we  have: 

RULE  110. — Divide  the  grate  area  in  square  inches  (as  found  by  Rule  103)  by  400; 
the  quotient  will  be  the  area  in  square  inches  of  each  orifice  in  a  double  exhaust  nozzle 
for  either  soft-  or  hard-coal  burners. 


508 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


Ill  this  way  we  have  found  the  areas  of  the  orifices  given  in  Column  3  in  the 
following  table,  and  in  Column  4  the  corresponding  diameters  are  given. 

TABLE  75. 

DIAMETERS   OP  ORIFICES   IN   DOUBLE  EXHAUST  NOZZLES   FOR   HARD  AND   SOFT  COAL  BURNERS 

AS   COMPUTED  BY  RULE   110. 


Size  of  Cylinders. 

Grate  Area. 

Area  of  each  Orifice. 

Diameter  of  each  Orifice. 

Column  1. 

Column  2. 

Column  3. 

Column  4. 

10  x  20  inches. 

954.40  sq.  in. 

2.38  sq.  in. 

1J  inches. 

11   x  22 

1267.20 

3.16        ' 

2        " 

12  x  22 

1425.60 

3.56        ' 

2i      " 

13  x  22 

1627.20 

4.06        ' 

2±      " 

14  x  24 

1843.20 

4.60        ' 

2ft    " 

15  x  24 

2116.80 

5.29        ' 

2ft    " 

16  x  24 

2275.20 

5.68       ' 

2H    " 

17  x  24 

2491.20 

6.22       ' 

m  " 

18  x  24 

2750.40 

6.87       " 

2U    " 

20  x  24 

4320.00 

10.80       " 

3ft    " 

22  x  24 

5011.20 

12.52      " 

4        " 

The  grate  areas  given  in  this  table  are  the  same  as  those  given  in  Tables  59  and  GO. 

The  size  of  orifices  in  single  exhaust  nozzles  or  thimbles  can  be  found  by  the 
following  rule : 

RULE  111. — Divide  the  grate  area  in  square  inches  (as  found  by  Rule  103)  by  200 ; 
the  quotient  will  be  the  area  in  square  inches  of  the  orifice  in  a  single  exhaust  nozzle 
or  thimble  for  either  hard-  or  soft-coal  burners. 

The  sizes  in  the  following  table  have  been  found  by  this  rule. 

TABLE  76. 

DIAMETERS  OF  ORIFICES   IN  SINGLE   EXHAUST  NOZZLES  OR  THIMBLES  FOR  HARD  AND  SOFT   COAL 

BURNERS  AS  COMPUTED  BY  RULE   111. 


Size  of  Cylinders. 

Grate  Area. 

Area  of  Orifice. 

Diameter  of  Orifice. 

Column  1. 

Column  9. 

Column  3. 

Column  4. 

10  x  20  inches. 

954.40  sq.  in. 

4.72  sq.  in. 

2|  inches. 

11   x  22 

1267.20 

6.33       " 

2J      " 

12  x  22 

1425.60 

7.12       ' 

3ft    " 

13  x  22 

1627.20 

8.13       ' 

3J      " 

14  x  24 

1843.20 

9.21        ' 

3ft    " 

15  x  24 

2116.80 

10.58       ' 

m  " 

16  x  24 

2275.20 

11.37        ' 

3{J    " 

17   x  24 

2491.20 

12.45       ' 

4        " 

18  x  24 

2750.40 

13.75       ' 

4ft    " 

20  x  24 

4320.00 

21.60        ' 

5i      " 

22   x   24 

5011.20 

25.05        ' 

5tJ     " 

The  grate  areas  given  in  this  table  are  the  same  as  those  given  in  Tables  59  and  60. 

The  diameters  of  orifices  as  computed  by  the  foregoing  rules  agree  very  closely 
with  the  diameters  in  actual  use,  as  will  be  seen  by  a  comparison  with  those  given  in 
the  following  table,  which  contains  the  sizes  of  orifices  as  found  in  locomotives  in 
regular  service. 


MODERN    LOCOMOTIVE    CONSTRUCTION. 


509 


111  Columns  2  and  3  are  given  the  sizes  of  three  nozzles,  comprising  a  set  for  each 
engine  such  as  are  generally  furnished  by  locomotive  builders.  Whenever  thimbles 
are  used  in  place  of  nozzles,  the  diameter  of  the  hole  in  top  of  the  exhaust  pipe  is  made 
£  inch  greater  than  the  diameter  of  largest  orifice  in  a  set  for  the  smaller  engines ;  and 
about  $  of  an  inch  greater  for  the  larger  engines. 

TABLE   77. 

SIZES   OF  ORIFICES   IN   EXHAUST  NOZZLE   IN   ACTUAL   SERVICE   FOR  HARD   AND  SOFT   COAL  BURNERS. 


Size  of  Cylinders. 

Double  Nozzles. 
Diameter  of  Orldcee  In  Inches. 

Single  Nozzles. 
Diameter  of  Orifices  In  inches. 

Column  1. 

Column  t. 

Column  3. 

9  x  16  inches. 

14,  if,  i*. 

12  x  20 

It,  1*,  24- 

13  x  20 

If,  2,     24. 

14  x  24 

11,  24,  2|. 

21,  34,  3f 

15  x  24 

2,     24,  24. 

34,  3|,  3f. 

16  x  24 

24,  24,  2}. 

3|,  3|,  3J. 

17  x  24 

24,  2J,  3. 

3»,  4,     44. 

18  x  24 

2fc  3,     34. 

4,     44,  44. 

19  x  24 

3,     34,  34. 

44,  4i,  5. 

20  x  24 

34,  34,  3f. 

4J,  5,     54. 

CHAPTER   XII. 


SAND-BOXES. BELLS. PILOTS. ENGINE-BRACES. 


SAND-BOXES. 


519.  The  sand-boxes  are  usually  placed  on  top  of  boilers;   sometimes  they  are 
placed  underneath  the  running  boards.     The  capacity  of  a  sand-box  placed  on  top  of 


Fig,  799 


Fig.  800 


boiler  is  generally  about  7  cubic  feet,  excepting  those  which  are  placed  on  small  engines, 
which  contain  less. 


MOI>I:I;\  I.OI-OMOTI  1 i-: 


511 


The  sand  should  be  perfectly  dry,  so  as  to  enable  it  to  flow  easily  through  the  pipes 
which  lead  it  to  the  rails.  It  should  be  used  sparingly,  not  simply  for  the  sake  of 
economy,  but  mainly  to  avoid  an  increase  of  train  resistance,  which  will  occur  when  the 
wheels  under  the  cars  have  to  run  on  sanded  rails ;  the  purpose  of  putting  sand  on  the 
rails  is  to  assist  the  engine  to  overcome  the  train  resistance ;  hence  the  latter  should 
not  be  increased  by  an  injudicious  use  of  sand. 

Fig.  797  shows  a  section,  Fig.  798,  an  elevation,  and  Fig.  799,  a  plan,  of  a  small 
sand-box  of  the  ordinary  form.  It  consists  of  a  cast-iron  base  A,  a  wrought-irou 
body  B,  generally  No.  12  B  W.  G.,  a  cast-iron  top  (7,  and  a  cover  D.  The  wrought- 
iron  body  B  is  not  riveted  to  the  base  or  top ;  the  body  is  simply  set  in  these  castings 
and  the  whole  clamped  together  by  the  rods //which  are  tapped  into  the  lugs  e  e  cast 
to  the  sand-box  top,  with  the  nuts 
on  the  lower  ends  of  the  rods  bear- 
ing against  the  underside  of  the  base. 
The  sand  is  shut  off  or  allowed  to 
escape  by  the  valves  h  and  i.  The 
valve  h  is  worked  directly  by  a  bell 
crank,  and  the  valve  i  is  worked  by 
the  bent  lever  /,  whose  free  end  is 
connected  by  a  link  k  to  one  of  the 
arms  of  the  bell  crank.  The  free 
end  of  the  arm  j  of  the  bell  crank  is 
connected  to  a  rod  extending  to  the 
cab,  within  easy  reach  of  the  engi- 
neer; this  rod,  or  sand-box  reach- 
rod,  as  it  is  often  called,  is  sometimes 
run  through  the  hand  rail.  It  will 

be  seen  that  both  valves  are  worked  simultaneously.  The  sand-pipes, 
which  are  usually  made  of  wrought-iron,  l£  inches  outside  diameter, 
are  fastened  to  the  lugs  m;  one  way  of  fastening  them  is  shown 
on  a  larger  scale  in  Fig.  800.  The  end  of  the  sand-pipe  S  is  turned 
over  on  a  flange  w,  and  this  flange  is  secured  to  the  lug  m  by  two 
studs  Fig.  801  shows  another  method  of  fastening  the  sand-pipe 
to  the  base.  The  lug  m  is  tapped,  a  nipple  is  screwed  into  it,  and  the  sand-pipe  is 
attached  to  the  nipple  by  a  nut,  plainly  shown  in  the  illustration. 

The  sand-box  (Fig.  798)  is  set  on  top  of  the  lagging,  and  secured  to  the  boiler  by 
two  studs  tapped  into  the  boiler  shell  and  passing  through  the  lugs  g  g,  which  are  cast 
to  the  base. 

Sometimes  the  sand-box  rests  directly  on  the  boiler  shell,  and  the  lagging  fitted 
around  it.  Some  master-mechanics  object  to  this  method,  as  sparks  are  liable  to  fly 
between  the  base  and  lagging  and  set  fire  to  the  latter. 

Figs.  802,  803  show  a  large  sand-box.  It  consists  of  a  cast-iron  base  A ;  a  cast- 
iron  lower  ring  B;  a  sheet-iron  body  f;  an  upper  ring  7);  a  top  J?;  and  a  cover  F. 
The  whole,  excepting  the  cover,  is  clamped  together  by  the  rods/ 

Of  course?  the  sand-boxes  must  bo  water-tight,  and  since  none  of  the  joints  are 


Fig.  801 


512 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


calked  it  is  necessary  to  adopt  such  form  of  joints  as  will  not  permit  the  water  to  flow 
into  the  box.  It  is  for  this  reason  that  the  body  C  is  made  to  overlap  the  upper  edge  of 
the  lower  ring,  as  shown  at  _p,  Fig.  802 ;  and  the  upper  ring  overlaps  the  top  of  the  body 
(7,  as  shown  at  r ;  the  joint  s  is  not  recommended,  because,  unless  it  is  extremely  well 
made,  it  will  allow  the  water  to  flow  into  the  box ;  the  joint  s  in  Fig.  804  is  a  better  one. 
The  manner  of  operating  the  valves 
is  plainly  shown  and  needs  no  fur- 
ther explanation.  The  reach-rod, 
which  extends  to  the  cab,  •  is  at- 
tached to  the  free  end  of  the  lever,/. 
Figs.  804,  805  show  a  sand-box 
arranged  for  four  sand-pipes,  and 
consequently  it  has  four  valves  ; 
these  are  connected  so  that  only 
two  can  be  operated  simultaneously 


Fig.' 803 


/  T    Fi&804    \J 

^ »-" 


Fig.  805 

Otherwise  the  construction  of  this  box  is  the  same  as 


with  one  lever  and  reach-rod, 
that  shown  in  Fig.  802. 

Figs.  806,  807  show  a  form  of  sand-box  which  in  late  years  has  been  extensively 
adopted.  It  is  copied  from  English  design,  and  its  advantages  are,  that  it  is  simple  in 
construction,  and  can  be  kept  clean  with  less  labor  than  other  types.  We  do  not  favor 


MODKKN  LOCOMOTIVE  COXSTIIUCTIOX. 


513 


the  joint  between  the  body  and  base,  as  this  joint  is  liable  to  let  water  flow  into  the 
box.  A  better  way  is  to  allow  the  body  to  overlap  the  top  of  the  base,  and  this  can 
easily  be  done  without  interfering  with  the  outer  form  of  the  box. 

520.  In  eight-wheeled  passenger  engines,  the  sand-pipes  are  placed  in  front  of 
the  main  drivers;  in  Mogul,  ten- wheeled,  and  consolidation  engines,  they  are  some- 
times placed  in  front  of  the  front  drivers;  in  others,  in  front  of  the  main  drivers, 

and  occasionally  on  both  sides  of  the  main 
drivers.  In  switching  engines  it  is  desirable 
to  have  the  sand-pipes  on  each  side  of  a 
pair  of  drivers,  and  in  such  cases  the  sand- 
boxes with  four  sand-pipes  are  often  used. 
The  latter  boxes  are  also  frequently  used 
on  other  engines. 


I 


FT 


I'       -y>v-" 

1,       ** 


BELLS. 


521.  Fig.  808  shows  a  locomotive  bell  A, 
with    its  cast-iron    stand    7?  and   cast-iron 


_L 


Pig.  807 


Pig.  W)8 


Fig.  800 


yoke  C.  Fig.  809  simply  shows  a  side  elevation  of  the  stand  and  yoke.  The  stand  is 
placed  on  top  of  boiler  and  fastened  to  the  shell  by  two  studs  tapped  into  the  shell. 
The  position  of  the  bell  is  generally  such  as  will  harmonize  with  and  give  a  pleasing 
appearance  to  the  whole. 

The  foot  a  on  the  stand  B  should  be  sufficiently  deep  to  bring  its  upper  surface 
about  J  inch  above  the  lagging,  which  should  be  fitted  closely  to  the  sides  of  the  foot, 
so  as  to  prevent  sparks  from  setting  fire  to  it. 

The  yoke  turns  on  wrought-iron  pivots  e;  these  are  driven  into  the  yoke,  and  held 
in  position  by  a  dowel  pin;  the  driven  part  of  the  pivot  is  made  either  st might  or 


514 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


tapered,  and  the  projecting  part  of  the  pivot  is  often  made  about  £  inch  larger 
in  diameter  than  the  driven  part.  The  bell  is  generally  worked  from  the  cab  by  a 
bell-rope  attached  to  the  handle  D;  sometimes,  by  an  automatic  bell-ringer  driven 
by  steam. 


CONSTKUCTION   OF  BELL. 


522.  The  size  of  the  bell  is  generally  designated  by  the  diameter  of  its  mouth ;  if 
this  is  12  or  18  inches  in  diameter,  the  bell  is  said  to  be  a  12-  or  18-inch  bell.     After 


this  diameter  has  been  established,  the  other  dimensions  of  the  bell,  excepting  its 
shank,  by  which  the  bell  is  fastened  to  yoke,  are  derived  from  this  diameter.  For 
large  locomotives  18-  or  19-inch  bells  are  generally  used,  and  for  the  smaller  engines 


MOVERS'  LOCOMOTIVE   CONSTRUCTION. 


515 


13-inch  bells  seem  to  bo  the  favorite  ones.  For  designing  a  bell  of  any  size,  the 
following  method  may  be  adopted  and  satisfactory  results  expected. 

Draw  the  diameter  /«,  Fig.  810,  of  the  mouth  of  the  bell;  also  draw  the  center 
line  C  C  perpendicular  to  fa.  Divide  the  diameter  fa  into  10  equal  parts;  each  part 
is  called  a  stroke ;  this  diameter,  divided  into  strokes,  forms  a  scale  from  which  the 
other  dimensions  of  the  bell  are  laid  off.  Hence,  for  the  sake  of  convenience,  draw  a 
line/a  equal  in  length  to  the  diameter  of  the  mouth  of  the  bell  on  a  separate  piece 
of  paper,  which  is  to  serve  as  a  scale,  divide  fa  into  10  equal  parts,  and  again  divide 
one  of  the  end  divisions  into  10  equal  parts,  and  thus  complete  the  scale.  Since 
one  of  these  parts  (a  stroke)  is  divided  into  10  equal  parts,  we  have  constructed  a 
decimal  scale. 

Parallel  to  the  center  line  C  C  draw  a  line  b ;  the  distance  between  b  and  C  C 
must  be  equal  to  2£  strokes ;  in  other  words,  the  distance  between  these  lines  must  be 
equal  to  one-half  of  C  a.  The  distance  from  b  to  C  C  represents  one-half  the  diameter 
of  the  crown,  consequently  the  diameter  of  the  crown  is  equal  to  one-half  the  diameter 
fa  of  the  mouth  of  the  bell.  From  the  point  a  as  a  center,  and  with  a  radius  equal  to 
8  strokes,  describe  a  short  arc  cutting  the  straight  line  6  in  the  point  8 ;  join  the  points 
a  and  8  by  a  straight  line ;  divide  this  line  into  8  equal  parts ;  each  part  will  then  be 
equal  to  a  stroke.  Mark  the  points  of  division  1,  2,  3,  etc.,  as  shown,  and  through 
these  points  draw  lines  perpendicular  to  a  8.  These  lines  are  ordinates  to  the  curve 
through  center  of  the  metal  of  the  bell,  and  must  be  made  equal  to  the  length  given 
in  the  following  table ;  these  lengths  must,  of  course,  be  measured  from  the  line  a  8, 
with  the  scale  of  strokes. 

TABLE  78. 


Length  of  ordinate 

tt        tt        it 

U              It              U 

It           II           It 

U              11               U 

11               11               U 

It           tt           It 

U              tl              U 

throng] 

u 
tt 
u 
11 

It 

i  point  

1 

=  0.41  sti 
=  0.86 
=  1.02 
=  1.00 
=  0.87 
=  0.66 
=  0.39 
=  0.09 

oke. 

t 

t 

tt 
i 
t 
i 

t 

2 

it 

3 

tt 

4 

u 

5 

It 

6 

It 

7 

II 

.  8 

Through  the  ends  of  these  ordinates  draw  a  curve,  which  will  be  the  curve 
through  the  center  of  the  metal.  From  the  points  in  which  the  ordinates  meet  this 
curve  as  centers  describe  circles  of  the  following  diameter : 


TABLE  79. 
Diameter  of  circle  on  ordinate ,  .  1  =  0.70  stroke. 


a 
u 
u 

It 
u 
It 


2  =  0.45 

3  =  0.33 

4  =  0.26 

5  =  0.22 

6  =  0.20 

7  =  0.19 

8  =  0.18 


Tangent  to  these  circles  draw  the  outer  and  inner  curves,  which  will  determine 
the  thickness  of  the  metal.    At  the  bottom  of  the  bell  describe  an  arc  from  a  center  on 


516 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


the  line  /  a,  so  that  the  arc  will  be  tangent  to  the  outer  curve  and  pass  through  the 
point  a.  Another  arc  with  about  the  same  radius  must  be  described  passing  through 
the  point  a,  and  tangent  to  the  inner  curve. 

The  crown  of  the  bell  is  sometimes  slightly  curved  as  shown,  sometimes  it  is 
flat ;  in  either  case  the  lines  which  represent  the  outer  and  inner  surfaces  of  the  crown 
are  drawn  tangent  to  the  circle  on  ordinate  8.  After  a  bell  has  been  drawn  full  size 
by  the  method  here  given,  the  drawing  can  be  measured  with  an  ordinary  rule,  and 
its  dimensions  in  inches  obtained. 

For  determining  the  diameter  and  length  of  the  shank  D,  good  judgment  is 
required.  For  an  18-inch  bell  the  shank  is  generally  from  2£  to  2§  inches  in  diameter, 
and  about  2j  to  2g  inches  long.  The  bolt  for  fastening  the  bell  to  the  yoke  Y  is  from 
|  to  1  inch  in  diameter.  The  length  of  the  clapper  should  be  such  that  an  arc  passing 
through  the  center  d  of  the  circle  on  ordinate  1,  and  described  from  the  center  g,  shall 
pass  through  the  center  of  the  ball  on  the  clapper.  The  weight  of  the  clapper  should 
be  from  -fo  to  -^  the  weight  of  the  bell;  the  heavier  weight  is  to  be  used  for  the 
smaller  bells. 

The  proportions  here  given  answer  well  for  an  ordinary  locomotive  bell.  Should 
it  be  desirable  to  use  a  greater  or  smaller  thickness  of  metal,  make  the  ordinates 
through  the  points  of  division  on  the  line  8  a  the  same  length  as  before ;  and  the 
diameters  of  the  circles  on  these  ordinates  are  found  by  the  following  table : 


TABLE   80. 
Diameter  of  circle  on  ordinate 1  =  1 


stroke. 


2  =  0.653 

3  =  0.474 

4  =  0.380 

5  =  0.327 

6  =  0.291 

7  =  0.279 

8  =  0.267 


If  it  is  now  desirable  to  make  the  thickness  of  metal  on  ordinate  1  equal  to,  say, 
1J  inches,  describe  a  circle  1^  inches  diameter;  the  diameters  of  the  circles  on  the 
other  ordinates  are  found  in  the  following  manner :  Multiply  l£  inch  by  0.653,  which 
is  the  diameter  given  in  Table  80  for  the  circle  on  ordinate  2 ;  for  the  circle  on  ordinate 
3  multiply  14  inches  by  0.474,  which  is  the  diameter  given  in  the  table  for  ordinate  3, 
and  so  on  in  succession  for  all  the  other  circles,  and  then  proceed  as  before.  In  a 
similar  way  we  find  the  diameters  of  the  circles  for  any  other  thickness,  either  more 
or  less  than  an  inch  on  ordinate  1. 

Increasing  the  thickness  of  the  metal  at  the  bottom  of  the  bell  will  make  the 
diameter  of  the  crown  a  little  larger  than  one-half  the  diameter  of  the  mouth,  but  this 
difference  will  not  affect  the  tone  of  the  bell  to  any  appreciable  extent. 

If  it  is  desirable  to  have  a  fillet  h  near  the  bottom  of  the  bell,  it  should  be  drawn 
tangent  to  the  circle  on  ordinate  1,  as  shown  in  Fig.  811.  A  good  mixture  for  the 
metal  is  4  of  copper  to  1  of  tin ;  to  every  hundred  pounds  of  this  mixture  add  £ 
pound  of  zinc,  and  £  pound  of  lead. 


MODERX  LOCOMOTIVE   CONSTRUCTION. 


517 


PILOTS. 

523.  Figs.  812,  813,  814  show  a  wooden  pilot  of  ordinary  design.  The  drawing  is 
very  complete,  and  therefore  a  few  remarks  only  are  necessary.  The  height  of  all 
pilots  should  be  such  as  to  leave  1  inch  clearance  between  the  top  of  rails  and  the 
bottom  of  the  pilot  when  the  engine  frames  rest  on  top  of  the  driving  boxes,  conse- 
quently when  the  engines  stand  at  their  ordinary  working  height  there  will  be  from 
3  to  5  inches  clearance  between  the  bottom  of  the  pilots  and  the  rails.  A  wrought-iron 

band  about  3  inches  wide  and  \ 
inch  thick  is  fastened  to  the 
sides  of  the  lower  side  rails  I 
of  the  pilot,  and  this  band  ex- 
tends partly  over  the  sides  of 
the  i%ear  rail  m.  The  pilot  is 
bolted  to  the  bumper  beam  C; 
its  front  end  is  braced  by  two 
rods  D,  which  run  up  to  the 
casting  E  fastened  to  the  top 
of  bumper  beam.  The  bottom 
of  the  rear  end  of  the  pilot  is 


Fig.  813. 


Fig.  814 


braced  by  the  braces  G,  which  cross  each  other  and  are  bolted  to  the  engine  frames, 
or  sometimes  to  the  cylinder  saddle.  For  further  security  the  cast-iron  knees  F  are 
used.  The  pulling  bar  A  is  attached  to  the  jaw  Z?,  whose  shank  passes  through  the 
bumper  beam  and  is  fastened  to  it. 

Figs.  815,  816,  817  show  another  wooden  pilot,  of  slightly  different  construction. 
Instead  of  bracing  its  front  end  by  two  braces,  only  one  brace  I)  is  used,  and  this  brace 
is  placed  inside  of  the  pilot  and  bolted  to  the  bottom  of  the  bumper  beam.  The  pulling 
bar  A  is  attached  to  a  cast-iron  draw-head  B.  The  draw-heads  are  made  of  various 
forms.  The  one  here  shown  has  its  upper  edge  flaring  forward ;  the  object  of  this 
flare  is  to  hold  obstructions  which  may  slide  along  the  top  of  the  pilot,  and  prevent 
them  from  smashing  the  front  end  of  the  smoke-box. 

Figs.  818,  819  show  a  wrought-iron  draw-head  which  is  sometimes  used  in  place  of 
the  cast-iron  head.  It  is  bolted  to  the  front  of  the  bumper  beam;  it  not  only  serves 


518 


MODERN  LOCOMOTIVE    CONSTRUCTION. 


as  a  draw-head,  but  also  as  a  bumper  bar.     The  brace  A  is,  of  course,  bolted  to  the  top 
of  bumper  beam. 

Figs.  820,  821  show  an  iron  pilot  with  two  wooden  bumper  blocks  B ;  these  blocks 


Fag.  815 


[!       '  4FT       ~5t± 

|* S3X^- 


Fig;  811 


are  also  used  occasionally  with  wood- 
en pilots. 

The  construction  of  this  iron 
pilot  is  clearly  shown,  and  needs  no 
further  description. 

BUMPER  BEAMS. 

524.  The  bumper  beams  are  gen- 
erally made  of  oak.     They  are  made 
long  enough  to  project  1  inch  beyond  the  cylinder  flanges ;  the  object  of  this  is  to  pre- 
vent obstructions  on  the  track  from  striking  the  cylinders. 

Sometimes  the  cross-section  of  the  bumper  beam  is  rectangular,  as  shown  in  Fig. 
813.  The  usual  size  of  this  cross  section  is  7  x  14  or  15  inches ;  frequently  the  cross- 
section  is  square  or  nearly  so ;  9  x  10  inches  is  often  used  for  engines  of  medium  size. 

The  bumper  beams  are  fastened  to  the  frames  in  various  ways ;  the  manner  of 
fastening  them  often  depends  on  the  distance  from  the  rail  to  the  draw-head  on  the 
cars. 

Fig.  822  shows  the  beam  bolted  to  the  end  a  of  the  frame  by  a  horizontal  bolt  c. 


LOCOJ/or/r£   CONSTRUCTION. 


519 


520  MODERN  LOCOMOTIVE   CONSTRUCTION. 

and  two  vertical  bolts  b  b,  which  also  pass  through  the  foot  of  the  bumper  brace  A ; 
this  foot  is  lipped  over  a  projection  on  the  frame,  so  as  to  reduce  as  much  as  possible 
the  stress  in  the  frame  F. 

Sometimes  a  casting  shown  in  Figs.  300,  301  is  used ;  a  wrought-iron  plate  about 
1  inch  thick  extending  across  the  engine  is  bolted  to  the  front  faces  of  these  castings 
and  the  beam  bolted  to  the  plate. 

525.  In  ordinary  service,  the  principal  stresses  in  the  beam  are  those  due  to  the 
pull  and  push  of  the  engine,  the  weight  of  the  pilot,  and  the  stress  due  to  removing 
comparatively  slight  obstructions. 

The  simplest  way  to  proceed  in  determining  the  size  of  cross-section  is  to  assume 
that  the  beam  has  to  resist  the  pull  and  push  of  the  engine  only,  and  adopt  for  this 
work  a  factor  of  safety  sufficiently  high  to  allow  for  the  other  stresses. 

In  calculating  the  strength  of  bumper  beams,  they  should  be  considered  as  beams 
supported  by  the  ends  of  the  frames  and  loaded  at  the  center.  When  a  draw-head  is 
used  as  shown  in  Fig.  816,  the  load  is  really  applied  at  two  points  when  the  engine  is 
pulling,  and  it  is  uniformly  distributed  over  a  short  distance  when  the  engine  is 
pushing;  therefore,  if  we  consider  the  beam  to  be  loaded  at  the  center  as  stated, 
we  shall  obtain  a  somewhat  larger  beam  than  necessary,  but  the  error  will  be  on 
the  side  of  safety;  we  shall  therefore  consider  all  bumper  beams  to  be  loaded  at 
the  center.  The  load  is  numerically  equal  to  the  adhesion  of  the  driving  wheels,  and 
the  direction  in  which  the  load  acts  is  of  course  horizontal  instead  of  vertical  as  in 
ordinary  beams. 

For  computing  the  width  or  breadth  of  beam  we  use  the  formula 

Sbd* 
W  —  n 


61 

in  which  W  denotes  the  load  in  pounds ;  the  value  of  n  depends  on  the  manner  of 
loading  and  supporting  the  beam — in  this  case  the  value  of  n  will  be  4 ;  8  denotes  the 
stress  per  square  inch  on  the  fiber  most  remote  from  the  neutral  surface,  and  for  a 
breaking  stress  a  value  of  14,400  maybe  given  to  S;  b  denotes  the  breadth;  d,  the 
depth ;  and  I,  the  length  of  the  beam.  All  the  dimensions  are  to  be  taken  in  inches. 
Substituting  these  values  for  the  corresponding  symbols  in  the  foregoing  formula,  we 

get, 

14400  x  b  x  d2 

w=±x-  -eirr 

This  reduces  to 

9600  x  I  x  d2 


W  = 


I 


The  above  gives  us  the  breaking  load ;  for  a  safe  load  which  a  beam  can  support,  and 
which  is  subjected  to  shocks  like  a  bumper  beam,  we  should  adopt  a  factor  of  safety  of 

9600 
15.     Therefore,  instead  of  using  a  multiplier  of  9600  we  should  use  ^—   =  640.     In 

order  to  allow  for  the  minor  stresses  to  which  the  bumper  beam  is  subjected,  we  must 


MODERN  LOCOMOTIVE   CONSTRUCTION.  521 

make  a  further  reduction,  say  to  480.    Hence,  for  the  safe  load  on  the  bumper  beam  we 
have  the  following  formula : 

w  =  ™x?xd';  (») 


or, 


x  I  =  480  x  b  x  d3.  (I) 


We  are  now  in  a  position  to  apply  the  formula. 

EXAMPLE  145. — Find  the  length  of  a  side  of  a  cross-section  of  a  square  bumper 
beam  for  an  eight-wheeled  passenger  engine  with  cylinders  17  inches  diameter  and  18 
inches  stroke ;  the  distance  between  the  centers  of  frames  is  47  inches. 

In  Table  5  we  find  that  the  adhesion  of  a  17  x  18  inch  eight-wheeled  passenger 
engine  is  10,404  pounds.  Now,  substituting  in  the  last  formula  for  the  symbols  their 

values,  we  have, 

10404  x  47  =  I  x  d-  x  480. 

Since  in  square  sections  I  is  equal  to  d,  we  have, 

10404  x  47  =  d3  x  480; 
hence, 

„_  10404^7 

and 


d  =  V  1018.72  =  10.1  inches  nearly. 

This  size  of  beam  will  answer  the  purpose  when  a  draw-head  as  shown  in  Fig.  815 
is  used,  or  in  which  no  large  hole  is  bored  through  the  center  or  dangerous  section. 
When  a  pulling  jaw,  as  shown  in  Fig.  812,  is  employed,  a  large  hole  must  be  bored 
through  the  center  of  the  beam,  thereby  greatly  reducing  its  strength,  which  must 
not  be  neglected.  To  show  how  formula  (b)  can  be  applied  in  this  case  we  will  take 
the  following  example : 

EXAMPLE  146. — The  cross-section  of  a  bumper  beam  for  an  eight-wheeled  passenger 
engine  with  17  x  18  inch  cylinders  is  to  be  rectangular,  as  shown  in  Fig.  813 ;  the 
breadth  is  7  inches;  diameter  of  hole  through  center,  2J  inches;  distance  between 
frames  47  inches.  What  must  be  the  depth  of  the  beam  ? 

The  hole  through  the  center  reduces  the  breadth  from  7  inches  to  7  —  2j  =  4£ 
inches,  and  this  value  must  be  taken  for  b  in  the  formula. 

Now,  substituting  in  formula  (b)  for  the  symbols  their  numerical  values,  we  have, 

10404  x  47  =  4.5  x  d°-  x  480; 
hence, 

10404  x  47   _  488988  _ 

4.5  x  480          2160  A38' 

and  

d  =  V226.38  =  15  inches. 

When  the  cross-section  of  a  bumper  beam  is  rectangular,  it  is  always  best  to 
choose  first  the  breadth  b,  which  in  many  cases  is  arbitrary,  and  for  choosing  this 
dimension  good  judgment  must  be  used.  The  depth  may  then  be  found  by  the  fore- 


522 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


going  formula,  from  which  the  following  rule  may  be  established  for  computing  the 
depth  of  beams  of  rectangular  cross-section. 

EULE  112. — Multiply  the  adhesion  in  pounds  on  the  drivers  by  the  distance  in 
inches  between  the  centers  of  the  engine  frames ;  call  this  product  A.  Multiply  the 
breadth  in  inches  of  the  bumper  beam  by  a  constant  of  480;  call  this  product  B. 
Divide  product  A  by  the  produ3t  B,  and  extract  the  square  root  from  the  quotient ; 
the  result  will  be  the  required  depth  in  inches  of  the  beam. 


FINISH  OF  FKONT  END. 

526.  Figs.  822,  823  show  the  general  finish  of  a  front  end  of  an  engine  with  a 
short  smoke-box.  It  consists  of  a  bumper  plate  K  covering  the  beam  and  extending 
backwards  up  to  the  cylinder  saddle.  Its  side  edges  are  curved  so  as 
to  allow  sufficient  room  for  taking  off  the  cylinder  heads  and  removing 
the  pistons  when  necessary.  Side  plates  riveted  to  the  bumper  plate 
extend  along  these  edges  and  along  the  front  of  the  bumper  beam. 

Sometimes  the  bumper  sheet  E  extends  only  from  outside  to 
outside  of  frames,  leaving  those  portions  of  the  beam  which  project 
beyond  the  frames  exposed ;  or  if  it  is  desirable  to  pro- 

_J+  \£\  *ect  these  portions,  they  are  covered  by  short  plates, 

and  the  ends  of  the  beam  are  covered  with  cast-iron 
caps. 

Fig.  822        U-u-'-^i 

BOILER  SUPPORTS  AND  FRAME  BRACES. 

527.  For  short  smoke-boxes  the  type  of  bumper 
brace  A  as  shown  in  Fig.  822  is  always  used.  Its 
upper  end  is  sometimes  made  rectangular,  as  shown, 
and  bolted — sometimes  riveted — to  the  smoke-box ;  the 
bolts  should  be  turned  and  driven  into  reamed  holes, 
two  of  the  bolts  passing  through  the  smoke-box 
ring.  Sometimes  the  upper  end  is  made  circular  in 
form,  with  one  bolt  passing  through  the  smoke-box 
ring. 
This  type  of  brace  is  occasionally  adopted  for  extension  fronts,  but  in  the 

majority  of  cases  the  type  shown  in  Figs.  824, 825  is  the  favorite 

one.     The  foot  a  of  this  brace  is  sometimes  set  on  top  of  and 

bolted  to  the  bumper  beam,  but  usually  it  is  placed  close  to 

the  rear  of  it  on  the  top  of  engine  frame,  and  bolted  to  the  lat- 
ter.    Its  upper  end  b  is  riveted  to  the  extension  front. 

528.  The  next  brace,  extending  from  the  engine  frames  to 

the  boiler,  is  the  guide-yoke.    These  braces  may  be  divided  into 

two  principal  classes:  one  class  embraces  those  which  do  not 

extend  from  frame  to  frame  and  consequently  a  separate  guide- 


Fig.  823 


Fig.  824 


yoke  is  required  for  each  side  of  the  engine ;  the  second  class  embraces  those  which 
extend  across  both  frames. 


MUDKKX   LOCOMOTIVE   CONSTRUCTION. 


523 


Figs.  826,  827  show  a  guide-yoke  which  does  not  extend  across  the  engine; 
it  lias  a  foot  a  forged  to  it,  which  is  bolted  to  the  frame;  the  holes  d  d  take  the 
rocker-box  bolts ;  this  box  also  serves  to  hold  the  guide-yoke  iu  position.  The  upper 
end  is  riveted  to  the  boiler  by  the  angle  iron  I.  The  holes  c  c  take  the  shanks 
of  the  guide  blocks.  The  opening  e  must  be  made  long  enough  to  clear  the 
connecting-rod.  This  brace  is  often  used  on  eight-wheeled  passenger  engines.  Its 
thickness  is  f  or  5  inch  in  large  engines,  and  a  thickness  of  {?  inch  is  sufficient  for 
smaller  ones.  Figs.  828,  829  show  the  second  kind  of  guide-yoke ;  it  is  generally  used 
on  consolidation  engines,  and  frequently  on  Mogul  and  ten-wheeled  engines.  Soine- 


Fig.  820 


Fig.  827 


times  the  feet  by  which  the  guide-yoke  is  fastened  to  the  frames  are  forged  to  it,  but 
we  believe  that  the  best  practice  is  to  use  cast-iron  brackets  a  bolted  to  it  and  to  the 
frames,  as  shown  in  the  figures. 

The  edge  c  must  be  of  sufficient  height  above  the  frames  to  clear  the  side-rods, 
At  6  the  brace  is  cut  out  to  clear  the  valve-rod.  The  piece  d  is  simply  a  guard  for  the 
connecting-rod.  The  hole/  takes  the  shank  of  the  guide  block. 

529.  A  wrought-irou  plate,  usually  called  the  belly-brace,  about  f  or  £  inch  thick, 
is  bolted  to  the  center  of  the  brace  at  e  e  e,  Fig.  828,  extending  to  the  bottom  of  the 
boiler.     This  plate  is  either  flanged,  or  has  an  angle  iron  riveted  to  it  to  fit  the  boiler. 
This  flange  or  angle  iron  should  not  be  fastened  to  the  boiler,  as  it  is  in  some  cases. 
The  object  of  the  belly-brace  is  simply  to  prevent  a  lateral  movement  of  the  frames, 
which  the  curvature  of  the  upper  flange  or  angle  iron  will  do  without  fastening  it  to 
the  boiler. 

530.  Fig.  830  shows  the  position  of  the  frame  braces  in  an  eight-wheeled  pas- 
senger engine.     The  brace  A  is  bolted  to  the  frame,  and  extends  to  the  boiler,  to  which 
it  is  riveted.     Separate  views  of  this  brace  are  shown  in  Figs.  831,  832.     The  brace 
7?  (Fig.  830)  extends  from  one  frame  to  the  other,  and  is  bolted  to  them  by  studs. 


524  MODERN  LOCOMOTIVE   CONSTRUCTION. 

Separate  views  of  this  brace  are  shown  in  Figs.  833,  834.  A  belly-brace,  shown  in 
Figs.  835,  836,  is  bolted  to  it.  The  holes  a  (Fig.  835)  take  the  rod  for  the  lifting-shaft 
spring  balance. 

The  brace  C  (Fig.  830)  is  an  ordinary  lip-brace ;  it  is  shown  separately  in  Fig.  837. 

531.  Fig.  838  shows  the  position  of  frame  braces  in  a  Mogul  engine.     These 
engines  have  a  two-wheeled  truck,  and  consequently  the  braces  A  A  in  rear  of  the 
bumper  beam  are  required  for  the  truck  center-pin  guide.    These  braces  are  shown  sepa- 
rately in  Fig.  839.     A  cast-iron  thimble  B  is  in  some  engines  placed  between  the  two 
frame  splices  8  and  T,  and  the  whole  bolted  together.     The  braces  C  C  simply  extend 
from  frame  to  frame,  and  have  a  belly-brace.     They  are  shown  separately  in  Fig.  840. 
D  D  represent  ordinary  lip-braces,  which  are  shown  separately  in  Fig.  837.     The  brace 
E  takes  the  pin  for  and  supports  the  radius  bar  of  the  truck;  separate  views  of  this 
brace  are  shown  in  Fig.  841. 

532.  Fig.  842  shows  the  position  of  frame  braces  in  a  consolidation  engine.     These 
braces  are  lettered  in  the  same  way  as  for  a  Mogul  frame,  Fig.  838.     Similar  letters 
indicate  the  same  kind  of  brace  in  the  two  engines,  so  that  what  has  been  said  in 
regard  to  the  braces  in  Mogul  engines  will  also  apply  to  those  in  the  consolidation 
engine. 

FOOT-PLATE  BRACES. 

533.  A  foot-plate  brace  7?  is  shown  in  Figs.  843,  844 ;  two  are  used  for  each  engine, 
They  are  generally  made  of  round-bar  iron,  varying  from  2  to  2£  inches  diameter- 
according  to  size  of  engine.     The  lower  foot  of  each  brace  rests  on  the  top  of  foot, 
plate  A,  and  is  secured  to  it  and  draw-bar  C  by  two  turned  bolts  driven  into  reamed 
holes.     The  upper  foot  is  usually  secured  to  the  boiler  head  by  four  studs.     It  will  be 
noticed  that  in  this  case  the  brace  is  rigidly  connected  to  the  boiler  and  foot-plate,  no 
attention  being  given  to  the  expansion  of  the  boiler ;  the  brace  is  supposed  to  spring 
enough  for  the  expansion.     The  advantage  claimed  for  this  rigid  connection  is  that  it 
tends  to  prevent  the  frames  from  being  permanently  injured  at  D  when  the  rear  end 
of  the  engine  has  to  be  hoisted  for  taking  out  the  driving  wheels,  because  when  this 
has  to  be  done  the  pedestal  cap  must  of  course  be  removed,  thereby  exposing  the 
weak  part  D  of  the  frames  to  .the  danger  of  bending.     In  some  cases  the  bolt  holes  in 
the  lower  foot  of  the  brace  are  made  oblong  to  allow  for  the  expansion,  but  from  what 
has  been  said  it  may  be  inferred,  and  correctly  too,  that  this  practice  is  liable  to  injure 
the  frame  at  D,  and  therefore  it  is  not  often  followed.     It  is  also  very  doubtful 
whether  the  oblong  holes  will  in  every  case  answer  their  purpose,  because  in  cases  of 
this  kind  the  braces  are  made  of  flat-iron  4  to  5  inches  wide  and  from  1  to  1J  inches 
thick,  which  makes  a  very  stiff  brace,  and  when  the  boiler  expands  it  is  liable  to  cant 
so  as  to  wedge  the  foot  between  the  bolt-heads  and  the  foot-plate,  thereby  rendering 
the  oblong  holes  useless  for  their  intended  purpose.     We  prefer  the  brace  and  the 
manner  of  fastening  it  as  shown  in  Fig.  843. 

In  some  engines  there  are  no  foot-plate  braces  used  •  in  such  cases  special  pro- 
visions are  made  for  hoisting  the  rear  end  of  engine. 


[fl 

I 

i 


i 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


527 


FOOT-PLATES. 

534.  Figs.  845,  846,  847,  848  show  different  views  of  a  cast-iron  foot-plate  A.  It 
reaches  from  the  rear  end  of  the  draw-bar  D  to  within  f  inch  of  the  boiler,  and 
finishes  flush  with  the  outside  of  the  frames,  with  the  exception  of  the  small  portions 
a  a  (Fig.  845),  which  are  cut  out  to  clear  the  spring  saddles.  A  part  h  of  the  rear  end 

lips  over  the  draw-bar  D,  and  forms  a  rubbing  plate.  A 
rib  k  butts  against  the  front  edge  of  the  draw-bar  D,  and 
the  ribs  m  m  bear  against  the  inner  sides  of  the  frames. 
The  ribs  1 1  extend  from  the  rib  k  to  the  pocket  «,  which 


ffig.846 


holds  the  pulling-bar  pin  shown  separately  in  Fig.  849.     The  pocket  n  should  always 
be  placed  as  far  forward  as  possible,  so  as  to  relieve  the  frames  of  lateral  stresses. 

The  spring  hangers  pass  through  the  slots  b  b  cut  through  foot-plate  and  frames ; 
the  ash-pan  damper  handles  pass  through  the  slots  c  c,  and  the  shaking  lever  works 


528 


MODERN    LOCOMOTIVE    CONSTKVCTION. 


in  the  slot  d.  The  holes  e  e  are  east  through  the  front  end  of  the  plate  for  the  feed- 
cock  rods,  which  are  required  only  when  pumps  are  used.  The  hub  g  (j  takes  the  end 
of  the  reverse  lever.  B  B  show  the  lower  ends  of  the  foot-plate  braces ;  C  C,  the 
safety  links  which  couple  to  the  safety  hooks  attached  to  the  tender ;  and  E  shows 
the  pulling-bar  support. 

Fig.  850  shows  the  wrought-iron  pulling-bar  which  connects  the  tender  to  the 
engine.  The  pin  holes  in  this  bar  are  bored;  one  of  them  takes  the  pin  shown  in 
Fig.  849,  and  the  other  one  takes  the  pin  through  the  draw-head  on  the  tender. 
These  pins  should  always  be  turned  and  made  to  fit  the  holes  through  the  pulling- 
bar  nicely.  The  distance  between  the  centers  of  these  holes  should  be  such  as  to 


'        -i 
-ffi—  —  ffl  —  ffl-  — 

fr  --  is*-"-  --  *  ----  -i2  —  £  ---  iax-  --- 


-  ---  -| 

~~ 


/taper  ^'lu  1 


allow  not  more  than  &  of  an  inch  play  between  rubbing-plate  h  on  the  engines  and 
that  on  the  tender.  The  diameters  of  the  pins  are  made  considerably  larger  than  is 
necessary  for  the  forces  which  they  have  to  resist ;  the  object  of  this  is  to  reduce  the 
wear,  and  prevent  as  much  as  possible  an  increase  of  the  play  between  the  engine  and 
tender  rubbing-plates. 

Figs.  851,  852,  853  show  another  foot-plate.  It  differs  in  design  from  the  fore- 
going principally  in  having  pockets  a  a  a  cast  in  it  for  the  safety  chains. 

Since  weight  at  the  rear  end  of  an  engine  is  in  many  cases  not  objectionable — in 
fact,  it  is  often  desirable— a  liberal  amount  of  metal  is  put  in  the  foot-plates ;  they  are 
much  stronger  than  they  need  to  be  for  the  forces  they  have  to  resist. 

A  wooden  floor  is  generally  fastened  to  the  top  of  the  foot-plates. 


MODERS  LOCOMOTirK   COXSTRVCTIOX. 


529 


DRAW-BARS. 

535.  Fig.  854  shows  two  views  of  a  draw-bar.  It  connects  the  ends  of  the  two 
frames,  and  for  this  purpose  it  is  notched  at  a  a  and  laid  in  the  recess  d  cut  in  the 
ends  of  frames  (see  Fig.  830) ;  the  I 

ends  of  frames  are  finished  flush 

b 

with  the  rear  edge  of  the  bar.  The      /£~~ 
ends  b  b  of  the  bar  project  beyond 
the  outer  sides  of  the  frames  and 
support  the  house  brackets,  of 
which  separate  views  are  shown     '  'jr 
in  Figs.  858  to  861.     This  form 


Jr  4- 


j 


•**• 


Pig  854. 


of  draw-bar  is  generally  used  in  eight-wheeled  passenger  engines,  or  engines  which 
have  a  foot-plate.  Sometimes  the  draw-bar  is  placed  in  a  vertical  position  and 
bolted  to  the  ends  of  the  frames.  The  projecting  ends  of  this  draw-bar  also  support 
the  house  brackets,  but  in  cases  of  this  kind  the  foot  of  the  house  bracket  is  formed 
as  shown  in  Figs.  864,  865. 

In  consolidation  engines,  or  other  classes  of  engines  in  which  hard  coal  is  burnt) 
the  boiler  extends  to  within  a  short  distance  of  the  ends  of  the  frames,  leaving  no  room 
for  a  cast-iron  or  other  kind  of  foot-plate.  In  these  engines  a  form  of  draw-bar  as 


Fig.  857 


shown  in  Fig.  855  is  used ;  Fig.  856  shows  a  plan  of  the  upper  bar,  and  Fig.  857  that 
of  the  lower  one.  These  bars  are  fitted  into  recesses  cut  in  the  ends  d  d  of  the  frames ; 
these  recesses  are  shown  at  e  e,  Fig.  842.  The  holes  a  a,  Figs.  855, 856,  take  the  pulliug- 
bar  pin,  and  the  holes  b  b  take  the  safety-chain  pins.  The  chaffing  block  c  is  bolted 
between  the  two  bars. 

HOUSE   BRACKETS. 

536.  Figs.  858,  859,  860,  861  show  different  views  of  a  house  bracket,  which  is  used 
in  connection  with  the  draw-bar  shown  in  Fig.  S54.  The  house  bracket  is  bolted  to  the 
draw-bar  d  by  the  shank  of  the  step  column  e  and  the  bolt  /;  it  is  also  bolted  to  the 
engine  frames/  the  foot-plate  a,  and  the  draw-bar  by  the  shank  of  the  feed-pipe  hanger 


530 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


h  and  the  bolt  m.  The  flange  i,  which  is  cast  to  the  bracket,  holds  the  side-sheet  which 
extends  forward  to  the  boiler,  and  whose  width  extends  from  the  top  of  foot-plate  to 
the  underside  of  the  running  board  b. 
A  longitudinal  section  of  this  side-sheet 
is  shown  at  a  a  a  in  Fig.  844. 

The  wheel  cover  of  the  form  shown 
in  Fig.  862  is  bolted  to  flange  k.  The 
house  or  cab  H  is  bolted  to  the  running 
board  I.  The  wrought-iron  cab  handle 


Plan  of  Top 

Fig.  861 


is  marked  c.     Another  form  of  wheel  cover  is  shown  in  Fig.  863.     It  is  usually  bolted 
to  the  frame  by  two  brackets,  and  sometimes  it  is  also  bolted  to  the  running  board. 


MODKKX  LOCOMOTIVE   CONSTRUCTION. 


531 


Some  master-mechanics  prefer  the  latter  because  it  prevents  the   mud  from  being 
thrown  on  the  working  parts  of  the  engine. 

Figs.  864,  865,  866  show  another  house  bracket ;  it  is  used  in  connection  with  draw- 
liars  placed  vertically,  as  mentioned  in  Art.  535.  The  cab  handle  c  is  cast  to  the 
bracket. 

KUNNING  BOARDS. 

537.  The  running  boards  are  fastened  to  the  sides  of  the  boiler.  They  are  made 
of  various  lengths  and  forms ;  in  a  few  cases  they  are  three  or  four  inches  longer  than 


^_" JjtlL-.44«i_I!_ £ ." '— -^tZ"         ff?" 

^  \         !         ^», 


Front  End 


"3.-y-_---=--^-         -.-----s;^ L--_--^.--.  -.^-j--S-^=£.-^^^-^-_^.--^^? 

~g6': — — ' |"     — a-'-  ,  >!• yp~  — -r" — 22? -p •&£ — •+•• — 


the  cab,  but  frequently  they  extend  from 
the  rear  end  of  the  cab  to  the  smoke- 
box.  In  ordinary  passenger  and  freight 
engines,  the  rear  end  of  the  running 
board  is  supported  by  the  house  bracket, 
and  along  the  side  of  the  boiler  it  is  sup- 
ported by  wrought-iron  brackets  placed 
from  28  to  36  inches  from  center  to 
center ;  separate  views  of  these  brackets 
are  shown  in  Fig.  872;  these  are  gen- 
erally riveted,  sometimes  bolted,  to  the 
boiler.  Running  boards  are  frequently 
made  of  wood  usually  2  inches  thick, 
with  their  outside  edges  faced  with 
sheet-brass,  or  iron  bands  or  angle  iron ; 
the  latter,  we  believe,  makes  the  best  job. 
Fig.  867  shows  a  wooden  running 
board.  Sometimes  they  are  made  of 
wrought-iron,  steel,  or  cast-iron.  Fig. 
870  shows  one  made  of  steel  finished 
off  with  an  angle  iron  riveted  to  its 
outer  edges. 


Fig.  863 


K gjj, J 


-fl-— _-  $Ht- 


In  regular  passenger  engines,  such  as  is  shown  in  Fig.  1,  the  engineer  stands  on 
the  foot-plate  ;  and  in  consolidation  engines  with  the  boiler  extending  to  the  draw- 


532 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


bar,  or  in  other  hard-coal  burning  engines  which  have  no  foot-plate,  the  engineer 
stands  on  the  running  board.  For  determining  the  distance  at  which  the  running 
board  is  to  be  placed  above  the  foot-plate,  good  judgment  must  be  used.  In  engines 
with  foot-plates  the  running  board  is  placed  at  a  convenient  height  to  serve  as  a  seat 


.-a-ix- 


-iS-8X- 


Fig.  i 


Fig.  868 


for  the  engineer  and  fireman.  Sometimes  this  height  is  determined  by  the  diameters  of 
the  driving  wheels ;  if  these  are  large,  the  running  board  is  placed  at  about  4  inches 
above  the  flanges  of  the  drivers,  and  in  many  passenger  engines  they  are  placed  an 
inch  or  two  above  the  reach-rod  which  connects  the  reversing  lever  to  the  lifting-shaft 


j 

•  —  \) 

1  Piece  Right 

"t 
f  \ 

H      i  £ 

A—                       No.8  Steel 

f 

•4V                                                                1 

^ 

1 

,, 

u" 

Pig.  870 


Fig.  871 


In  hard-coal  burning  engines,  or  those  which  have  no  foot-plate,  the  running  board  is 
placed  at  a  height  which  will  enable  the  engineer  to  reach  any  of  the  cocks  or  valves 
placed  on  top  of  the  boiler. 

In  some  classes  of  engines,  particularly  switching  engines,  no  house  brackets  are 
used,  the  cab  being  placed  directly  on  the  foot-plate.  In  such  cases  the  position  of  the 
running  board  is  a  matter  of  choice  and  good  judgment. 

Figs.  871,  872  show  the  position  of  a  running  board  for  an  engine  with  the  cab 
placed  on  the  foot-plate. 


MODEJtX  LOCOMOTIVE  CONSTRUCTION, 


533 


CABS. 

538.  Figs.  873,  874  show  different  views  of  an  ordinary  cab.  Its  height  often 
depends  on  the  position  of  the  running  board.  For  engines  in  which  the  engineer 
stands  on  the  foot-plate,  the  minimum  clear  height  above  the  foot-plate  should  not  be 


Fig.  873 


less  than  6  feet ;  and  for  engines  in  which  the  engineer  stands  on  the  running  board 
the  minimum  height  should  bo  6  feet  above  the  latter. 

In  engines  with  foot-plates  the  distance  from  the  roar  boiler  head  to  the  front  cud 
of  the  cab  varies  from  1'2  to  15  inches,  so  as  to  allow  sufficient  room  for  the  valves  and 


534  MODERN  LOCOMOTIVE   CONSTRUCTION. 

cocks  which  are  attached  to  the  boiler;  and  the  cab  is  made  wide  enough  to  allow 
room  in  the  front  end  for  a  door  A  at  each  side  of  the  boiler ;  these  doors  should  be  at 
least  14  inches  wide. 

In  engines  without  foot-plates  the  width  between  the  sides  of  the  boiler  and  the 
inside  of  cab  should  not  be  less  than  20  inches ;  24  inches  is  better,  if  the  width  of  the 
road  will  allow  it.  In  nearly  all  engines  the  length  of  the  cab  depends  on  the  length  of 
arc  described  by  the  end  of  the  reverse-lever  handle;  usually  from  9  to  12  inches 
between  the  extreme  positions  of  the  reverse-lever  handle  and  ends  of  the  cab  is 
allowed,  making  the  cab  from  5  feet  6  inches  to  6  feet  long. 

A  space  of  about  2  inches  is  usually  left  between  the  woodwork  in  the  front  end 
and  the  boiler  shell,  and  this  opening  is  closed  by  wrought-iron  plates  b  fitted  closely 
to  the  boiler  and  bolted  to  the  woodwork ;  they  serve  as  braces  for  holding  the  cab. 
These  plates  are  also  sometimes  bolted  to  the  shell  of  the  boiler  by  angle  plates  or 
brackets.  In  consolidation  or  hard-coal  burning  engines,  doors  are  required  in  the  rear 
end  of  the  cab.  In  this  class  of  engines  the  space  between  the  rear  end  of  the  cab  and 
the  boiler  head  is  closed  by  wrought-irou  plates,  leaving  the  furnace  door  exposed  so 
that  the  boiler  can  be  fired  from  the  tender.  The  front  windows  in  the  sides  of  the 
cab  are  generally  fixed  in  position,  and  the  rear  ones  are  made  to  slide  for  the  conven- 
ience of  looking  out. 


6H- 


-3K 


Fig.  879 


CHAPTER   XIII. 

ENGINE  TRUCKS. 

539.  Figs.  875,  876,  877,  878  show  different  views  of  the  ordinary  four-wheeled 
truck.  The  axle-boxes  A  work  in  cast-iron  pedestals  7?,  which  are  bolted  to  the 
wrought-iron  frame  C,  Sometimes  these  pedestals  are  made  of  wrought-iron ;  the 
general  form  of  these  is  shown  in  Fig.  879.  The  frame  is  supported  by  two  springs 
I)  D  (one  on  each  side),  and  these  are  attached  by  their  hangers 
to  the  equalizing  bars  E,  one  on  each  side  of  the  spring. 
The  ends  of  these  bars  are  turned  downwards  and  rest  in 
pockets  cast  in  the  upper  end  of  the  axle-box.  We  believe  the 
better  way  is  to  make  the  ends  of  these  bars  perfectly  plain 
and  allow  them  to  rest  on  a  convex  surface  of  the  box,  as 
shown  in  Fig.  881.  A  longitudinal  brace  F  (Fig.  876)  on  each 
side  of  the  truck  connects  the  bottoms  of  the  pedestals.  The 
center  swing  casting  7  is  supported  by  four  hangers  which 
swing  on  the  pins  77,  and  these  are  supported  by  the  transverse 
bars  G  G  bolted  to  the  top  of  frame.  The  lips  a  a  cast  to  the 
center  casting  7  are  for  the  purpose  of  holding  up  the  front  end  of  the  engine  in  case 
a  hanger  has  been  broken.  The  annular  groove  b  on  top  of  the  casting  receives  the 
center  pin  which  is  bolted  to  the  cylinder  saddles.  The  outer  diameter  of  this  groove 
should  not  be  much  greater  than  11  inches,  because  by  increasing  this  diameter  the 
leverage  with  which  the  friction  between  the  pin  and  casting  acts  will  also  be  increased, 
and  may  prevent  the  truck  from  turning  as  freely  as  it  should  do  in  running  over  a 
curve,  and  may  cause  it  to  run  off  the  track. 

The  transverse  bars  G  are  stiffened  sideways  by  the  longitudinal  braces  K  K. 

Two  shackles  L  for  the  safety  chain  are  generally  bolted  to  the  front  end  of 
the  frame,  the  other  ends  of  the  chains  are  attached  to  the  bumper  or  to  the  engine 
frames. 

Figs.  880,  881,  882,  883  show  different  views  of  another  four-wheeled  engine  truck. 
The  principal  difference  between  the  foregoing  truck  and  this  one  is  that  in  the  latter 
the  swing  center  casting  7  is  supported  by  a  cast-iron  center  plate  M  instead  of 
wrought-iron  transverse  bars;  otherwise  the  designs  are  similar,  so  that  a  further 
description  of  this  truck  is  unnecessary. 

Figs.  880,  881  show  the  wheel  covers  attached  to  the  truck  frame,  and  although 
this  is  good  practice,  it  is  by  no  means  a  universal  one.  In  many  cases  the  wheel 


536 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


UODERX  LOCOMOTIVE   CONSTRUCTION. 


537 


538 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


covers  are  fastened  to  the  engine  frames  and  are  made  much  wider  than  shown  in 
Fig.  880,  so  that  they  will  always  cover  the  wheels  when  running  over  a  curve. 

On  some  roads,  swing  center  castings  are  not  used ;  in  such  cases  a  center  plate  as 
shown  in  Figs.  884,  885  is  often  adopted ;  the  annular  groove  I  takes  the  center  pin  as 
before. 

Fig.  886A,  886B,  and  886C  show  different  views  of  another  four-wheeled  truck. 
This  truck  was  designed  by  the  Pennsylvania  E.  E.,  and  is  used  under  their  Standard 
passenger  engines  illustrated  in  Fig.  998.  The  construction  is  so  plainly  shown  that 
further  explanation  is  not  necessary. 

540.  The  distance  from  center  to  center  of  axles  must  be  such  as  to  prevent  the 
wheels  from  striking  the  cylinder  heads  when  the  engine  is  running  over  a  curve.  It 
is  found  that  68  to  70  inches  between  the  centers  of  axles  is  sufficient  for  all  ordinary 
curves  and  for  a  piston  stroke  not  exceeding  24  inches. 

The  equalizing  bars  E  E  (see  Fig.  876)  are  generally  made  1  inch  thick,  hence  the 
usual  problem  is  to  find  the  depth  of  the  bar.  When  bars  are  symmetrically  formed 
and  symmetrically  loaded,  we  need  to  consider  only  one-half  of  the  bar,  as  shown  in 
Fig.  886. 

To  illustrate  let  us  take  the  following : 

EXAMPLE  147. — Find  the  depth  of  the  equalizing  bar  E  (Fig.  876),  1  inch  thick,  for 
an  eight-wheeled  passenger  engine  with  cylinders  18  inches  diameter  and  24  inches 
stroke. 

Eeferring  to  Table  5,  we  find  that  the  total  weight  on  the  truck  is  28,680  pounds. 
This  includes  the  weight  of  the  truck ;  hence,  to  find  the  load  which  the  equalizing  bars 


n 


tM'lbl. 


n 

V. 

) 

( 

)                        ___„ 

% 

„ 

£ 

™                                          £ 

c 

1 

9185  ll)i. 

Fig.  886 


have  to  support,  we  must  subtract  the  weight  of  the  wheels,  axles,  and  axle-boxes  from 
the  total  weight.  Let  us  assume  that  the  weight  of  the  wheels,  etc.,  is  3,200  pounds  ; 
under  these  conditions  the  load  which  the  four  equalizing  bars  will  have  to  support  is 

25480 
28680  -  3200  =  25480  pounds  ;  and  the  load  on  each  bar  will  be  —  —     =  6370  pounds. 


One-half  of  this  load,  —  ~~  —  3185  pounds,  is  supported  by  each  spring-hanger  pin 

(Fig.  886)  at  18  inches  from  the  center  of  the  bar.  The  reaction  of  each  axle-box  is 
also  equal  to  3,185  pounds.  All  these  conditions  are  represented  in  Fig.  886,  which 
shows  one-half  of  the  bar.  At  B,  18  inches  from  the  center  C,  equal  to  one-half  the 
distance  between  the  spring-hanger  pins,  we  have  a  force  of  3,185  pounds  acting  ver- 


MUI>/-:I;\ 


COXSTRUCTIOX. 


540  MODERN  LOCOMOTIVE   CONSTRUCTION. 

tically  downwards ;  at  A,  35  inches  from  (7,  equal  to  one-half  the  distance  between  the 
centers  of  the  axles,  we  have  a  force  of  3,185  pounds  acting  vertically  upwards.  The 
force  B  tends  to  turn  the  lever  downwards  about  some  point  in  C,  and  the  force  A 
tends  to  turn  the  lever  in  the  opposite  direction.  The  effects  of  these  forces  in  produc- 
ing rotation  are  measured  by  their  moments.  According  to  Art.  256  the  moment  of 
the  force  B  is  equal  to  3185  x  1.5  =  4777.5  foot-pounds,  and  the  moment  of  the  force  A 
is  equal  to  3185  x  2.91  =  9268.35  foot-pounds.  Subtracting  the  moments  of  B  from 
that  of  A,  we  have  9268.35  —  4777.5  =  4490.85  foot-pounds,  and  this  is  bending  moment 
to  which  the  lever  in  Fig.  886  is  subjected ;  let  us  designate  it  by  M.  The  resisting 
moment  is  equal  to  b  x  </2  x  S,  in  which  b  is  the  breadth  in  inches  of  the  bar ;  d,  its 
depth  in  inches ;  and  8,  a  constant  multiplier  depending  on  the  kind  and  quality  of 
the  material.  For  wrought-iron  we  may  adopt  200.  Now,  for  equilibrium  the  bending 
moment  M  must  be  equal  to  resisting  moment ;  this  condition  is  represented  by  the 
following  formula : 

M  =  I  x  d2  x  8. 

Now  replacing  the  symbols  by  their  values,  we  have 

4490.85  =  1  x  d2  x  200, 
from  which  we  get 

4490.85 

-200^  =  22.4542  =  d«, 

and 

d  =  ^22.4542  -  4.73  inches. 

Here  we  have  made  no  allowance  for  the  hole  which  must  be  drilled  through  the 
bar  at  the  section  B,  nor  have  we  taken  the  weight  of  the  bar  into  account ;  but  these 
will  not  affect  the  result  much — indeed,  if  we  make  the  bar  4$  inches  deep  it  will  have 
sufficient  strength  to  resist  the  load.  For  a  more  precise  method  of  finding  the  resist- 
ance of  materials,  we  must  refer  the  reader  to  the  "  Text-Book  of  the  Mechanics  of 
Materials  and  of  Beams,  Columns,  and  Shafts,"  by  Professor  Mansfield  Merriman,  one 
of  the  best  books  treating  on  these  subjects. 


TWO-WHEELED   TEUCK. 

541.  Figs.  887,  888  show  -a  two-wheeled  truck,  which  is  sometimes  called  a  pony 
truck.  This  truck  was  designed  for  an  engine  with  cylinders  14  inches  diameter.  The 
weight  of  the  front  end  of  the  engine  is  sustained  by  the  equalizing  bar  A,  which 
works  on  the  fulcrum  pin  R,  and  this  in  turn  is  held  by  the  casting  T  bolted  to 
the  cylinder  saddle. 

Frequently  two  holes  about  5  inches  from  center  to  center  are  drilled  for  the  pin 
R  through  the  lever  A  and  casting  T;  the  pin  is  placed  in  either  one  of  these  holes, 
whereby  the  weight  on  the  truck  can  to  some  extent  be  adjusted  to  suit  the  require- 
ments of  the  engine.  In  some  engines  the  fulcrum  pin  R  is  not  used ;  the  upper  face 
of  the  equalizing  bar  A  rests  against  a  convex  surface  formed  in  the  cylinder  saddles 
or  a  pocket  bolted  to  the  saddles.  But  when  a  fulcrum  pin  R  is  used  we  see  no 
reason  for  placing  it  out  of  the  center  (vertically),  as  is  often  done,  and  as  is  shown  in 


COXSTRFCTIOX. 


541 


Fig.  887 ;  indeed,  there  is  an  objection  to  this,  as  it  makes  the  bar  weaker  than  when 
the  pin  is  placed  in  the  center. 

The  rear  end  of  the  equalizing  bar  A  is  connected  by  a  link  B  (see  Fig.  887)  to  a 
transverse  equalizing  bar  (7,  which  is  connected  to  the  driving-wheel  spring  hangers 


D  D.  The  front  end  of  the  equalizing  bar  A  is  connected  to  the  king-bolt  E,  which 
transmits  the  pressure  on  to  the  cylindrical  rubber  spring  /•';  sometimes  a  steel  spinil 
spring,  .-is  shown  at  /«'  in  Fig.  s«»l,  is  used  in  place  of  the  rubber.  The  spring  rests 
on  the  center  pin  G  (Fig.  SS7),  which  is  supported  by  tin-  swing  center  casting  //. 


542 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


.Fig.  890 


This  center  pin  G  slides  up  and  down  in  the  cast-iron  guide  which  is  bolted  to  the 
transverse  engine  braces  S  S. 

The  hangers  I  I  connect  the  swing  center  casting  to  the  cast-iron  frame  J}  of 
which  separate  views  are  shown  in  Figs.  894,  895.  This  frame  is  supported  by  the 

truck  springs  K  K,  which 
rest  on  the  spring  saddles, 
and  these  are  in  turn  sup- 
ported by  the  axle-boxes. 

The  cast-iron  pedestals  M 
are  bolted  to  the  frame  J. 
The  rear  end  of  the  radius 
bar  N  is  supported  by  the 
transverse  brace  0,  which  is 
bolted  to  the  engine  frames ; 
the  brace  0  also  holds  the 
pin  Y  around  which  the  radius  bar  swings ;  the  front  ends  of  the  latter  are  bolted  to 
the  frame  J,  and  an  oblique  brace  P  on  each  side  of  the  truck  connects  the  pedestals 
to  the  radius  bar. 

Figs.  891,  892,  893  show  another  two- wheeled  truck  designed  for  a  consolidation 
engine  with  cylinders  20  inches  diameter.  The  construction  of  this  truck  is  similar 
to  that  of  the  one  just  described,  with  the  exception  that  a  wrought-iron  frame  «7of 
rectangular  form  is  used  in  place  of  a  cast-iron  one ;  the  transverse  braces  L  L  bolted 
to  this  frame  support  the  swing  center  casting  as  before. 

Figs.  897,  898,  899,  900  show  another  two-wheeled  truck  of  similar  construction 
to  the  foregoing,  with  this  exception,  that  double  spiral  springs  instead  of  elliptical 
springs  are  used.  This  truck  was  designed  for  a  consolidation  engine  with  20  x  24 
inch  cylinder. 

542.  The  length  of  a  two-wheeled  truck  is  the  horizontal  distance  from  the  center 
of  the  axle  to  the  center  of  the  pin  Y  in  the  radius  bar  N  (see  Fig.  887).  To  determine 
this  distance  we  must  first  fix  the  position  for  the  truck  wheels ;  these  should  be 
placed  sufficiently  ahead  so  that  they  cannot  strike  the  cylinder  heads  when  running 
over  a  curve.  A  distance  X  (Fig.  888),  equal  to  6  inches  from  the  cylinder  heads  to  the 
side  of  the  wheels  when  the  engine  stands  on  a  straight  track,  is  generally  ample  for 
any  curve  over  which  the  engine  may  have  to  run.  Having  established  the  position 
of  the  truck  wheels,  we  have  also  established  the  distance  from  these  to  the  center  of 
the  rigid  wheel  base.  We  mention  this  because  this  distance  is  of  importance  in  the 
following  graphical  construction  for  finding  the  length  of  truck,  which,  as  we  have 
seen,  is  equal  to  the  horizontal  distance  from  the  center  of  the  axle  to  the  center  of 
the  pin  Y,  Fig.  887,  around  which  the  radius  bar  swings. 

To  find  the  length  of  the  truck,  draw,  as  in  Fig.  901A,  two  lines  0  Y  and  0  X 
perpendicular  to  each  other,  meeting  in  the  point  0 ;  make  0  Y  and  0  X  equal  to  the 
distance  from  the  center  of  the  rigid  wheel  base  to  the  center  of  the  truck  wheels. 
On  the  line  0  X  lay  off  a  point  A ;  the  distance  from  A  to  0  must  be  equal  to  one- 
half  of  the  rigid  wheel  base,  Join  the  points  A  and  Fby  a  straight  line  A  Y,  bisect 
this  line  by  a  perpendicular  B  C,  cutting  0  Y  in  the  point  B ;  the  distance  0  B  will  be 


544 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


the  length  of  the  two-wheeled  truck.  It  will  be  noticed  that  in  this  construction  we 
have  not  taken  the  radius  of  the  curve  into  account,  and  indeed  we  have  a  perfect 
right  to  omit  it,  because  the  radius  of  the  curve  will  not  affect  the  length  of  the  truck ; 
if  it  is  correct  for  one  radius,  it  will  also  be  correct  for  any  other  radius.  This  con- 
struction holds  good  for  all  classes  of  engines  having  a  two- wheeled  truck;  in  all 
these  engines  the  rigid  wheel  base  is  equal  to  the  distance  between  the  centers  of 


Fig.  901A 


front  and  rear  pair  of  drivers.     The  length  of  the  two-wheeled  truck  can  also  be  found 
by  the  following  computation : 

Let  a  in  Fig.  901A  designate  the  rigid  wheel  base;  6,  the  distance  from  the 
center  of  the  front  driving  axle  to  the  center  of  the  truck  axle ;  and  x,  the  required 

length  of  engine  truck.     Then  we  have 

I  (a  +  I) 

00  ' — '  ) 

which  reads : 

RULE  113. — Multiply  the  total  wheel  base  by  the  distance  from  the  center  of  the 
front  driving  axle  to  the  center  of  the  truck  axle,  and  divide  this  product  by  the  sum 
of  the  rigid  wheel  base  and  twice  the  distance  from  the  center  of  the  front  driving 
axle  to  the  center  of  the  truck  axle ;  the  quotient  will  be  the  length  of  the  truck. 

EXAMPLE  148. — The  rigid  wheel  base  of  a  consolidation  engine  is  13  feet  8  inches, 
and  the  distance  from  the  c°nter  of  front  driving  axle  to  the  center  of  truck  axle  is  7 
feet  10  inches ;  find  the  length  of  truck. 

Substituting  for  symbols  in  the  foregoing  formula  their  values,  we  have 

7'  10"  x  (13'  8"  +  7'  10^) 

13'  8"  +  15'  8" 
or,  reducing  to  inches,  we  have 

_  94  x  (164j+  94) 
x  ~~         164  +  188          : 


:i;S  LOCOMOTIVE   CONSTRUCTION. 


545 


546  MODERN  LOCOMOTIVE   CONSTRICTION. 

for  the  length  of  truck.  This  is  the  theoretical  value.  In  practice  it  is  customary  to 
deduct  10  per  cent,  so  as  to  insure  against  running  off  the  track  when  the  engine 
is  running  over  a  curve;  this  will  give  us  a  length  of  68.9  —  6.89  =  62.01,  say  62 
inches. 

ENGINE-TRUCK   EQUALIZING   LEVER. 

543.  The  length  of  the  equalizing  lever  A  (Fig.  887)  is  of  course  equal  to  the 
distance  from  the  center  of  the  truck  axle  to  the  front  spring  hangers  D  of  the  front 
driving-wheel  springs.  We  have  already  seen  that  the  weight  of  the  front  end  of  the 
engine  is  carried  by  this  equalizing  lever ;  and  it  will  readily  be  pei'ceived  that  a 
portion  of  this  weight  is  transmitted  to  the  front  hangers  of  the  front  driving-wheel 
springs,  and  the  other  portion  is  transmitted  to  the  king-bolt  through  the  truck  center 
pin.  If  the  fulcrum  R  (Fig.  889)  is  in  the  center  of  the  length  of  the  lever,  then  the 
weight  transmitted  to  the  truck  will  be  equal  to  that  transmitted  to  the  front  driving- 
wheel  spring  hanger.  But  these  springs  act  like  a  lever  with  its  fulcrum  at  the 
center,  hence  the  weight  on  the  driving  axle  will  in  this  case  be  equal  to  twice  that 
transmitted  to  the  spring  hangers.  But  in  many  cases  the  fulcrum  R  is  not  in  the 
center  of  the  length  of  the  equalizing  lever,  and  in  practice  we  usually  know  only 
the  weight  on  the  ti-uck.  Hence  the  following  practical  problem  often  presents  itself : 

The  weight  on  the  truck  being  known,  and  the  position  of  the  fulcrum  R  being 
established,  it  is  required  to  find  the  weight  transmitted  to  front  driving-wheel  springs, 
so  that  the  size  of  the  latter  can  be  computed.  To  show  how  problems  of  this  kind 
can  be  solved,  we  shall  give  an  example. 

EXAMPLE  149. — Let  the  length  of  the  equalizing  lever  (Fig.  889)  between  the 
centers  of  holes  A  and  B  be  80  inches ;  the  distance  from  center  of  hole  B  to  the  center 
of  fulcrum  R  be  43  inches ;  then  the  distance  from  A  to  R  will  be  37  inches ;  let  the 
whole  weight  on  the  truck  wheels  be  14,400  pounds.  It  is  required  to  find  the  weight 
transmitted  to  the  front  driving-wheel  springs. 

In  order  to  find  the  force  acting  on  the  end  A  of  the  lever,  we  must  subtract  the 
weight  of  the  truck  wheels,  axles,  boxes,  springs,  hangers,  and  spring  saddle  from  the 
total  weight  on  the  truck.  Assume  that  the  weight  to  be  deducted  is  1,900  pounds ; 
then  the  force  acting  at  A  of  the  lever  will  be  14400  —  1900  =  12500  pounds.  Now, 
by  the  principle  of  the  lever  we  know  that  the  product  of  the  distance  from  A  to  R 
into  the  force  acting  at  A  must  be  equal  to  the  product  of  the  distance  from  B  to  R 
into  the  force  acting  at  B.  Hence  we  have 

12500  x  37  =  43  x  force  at  B, 
and 

12500  x  37 

— r^ —    -  =  10755.8  pounds  for  the  force  acting  at  B. 

But  there  are  two  driving-wheel  springs  to  which  this  weight  is  transmitted,  hence  on 

10755.8 
the  end  of  each  spring  we  have  a  force  acting  vertically  which  is  equal  to  — 5— 

5377.9  pounds,  and  at  the  center  of  each  spring  we  have  a  force  tending  to  straighten 
the  spring  equal  to  5377.9  x  2  =  10755.8  pounds,  and  the  spring  must  be  made  strong 
enough  to  resist  this  force  or  load, 


MODERN  LOCOMOTIVE  CONSTRUCTION.  547 

544.  At  I)  (Fig.  890)  the  thickness  of  the  equalizing  lever  is  increased  and  the 
holes  R  R,  are  bushed  with  case-hardened  wrought-iron  thimbles  so  as  to  reduce  the 
wear  as  much  as  possible. 

To  find  the  depth  of  the  equalizing  lever  when  its  thickness  has  been  established ; 
or  to  find  the  thickness  of  the  lever  when  its  depth  has  been  established,  are  problems 
which  often  present  themselves  to  the  designer ;  their  solutions  are  readily  found  as 
follows : 

Take,  for  example,  the  lever  shown  in  Fig.  889.  Its  length  from  A  to  B  is  80  inches ; 
the  force  acting  at  its  end  A  is  12,500  pounds ;  the  fulcrum  pin  is  at  R,  37  inches  from 
A ;  find  the  thickness  and  depth  of  the  lever.  One  of  these  dimensions  is  first  arbi- 
trarily chosen ;  suppose  that  we  decide  to  make  its  thickness  2J  inches,  we  have  now 
to  find  simply  its  depth. 

The  dangerous  section  is  on  the  vertical  line  through  the  center  of  R,  and  the 
bending  moment  for  this  section  is  numerically  equal  to  the  product  of  the  distance 
A  R  into  the  force  acting  at  A ;  hence  we  have  for  the  bending  moment 

37  x  12500 

,n         =  38541.66  foot-pounds. 

Let  us  designate  this  bending  moment  by  M,  and  let  the  depth  in  inches  be  desig- 
nated by  d,  and  the  breadth  in  inches  by  b.  The  resisting  moment  will  then  be  equal 
to  d2  x  b  x  200.  But  the  bending  moment  is  equal  to  the  resisting  moment,  and  this 
condition  expressed  in  symbols  is 

M  -  d*  x  b  x  200. 
Substituting  for  the  symbols  their  values,  we  have 

38541.66  =  d2  x  2.25  x  200; 
hence 

38541.66 

rf2        -WO"    =85'64' 
and 

d  =••  V35M  =  9.25  inches. 

If  we  had  chosen  arbitrarily  the  depth  instead  of  the  thickness  of  the  lever,  its  thick- 
ness could  have  been  found  in  a  similar  manner.  For  example,  find  the  thickness  of 
the  equalizing  lever  shown  in  Fig.  889 ;  it  is  80  inches  long,  its  depth  at  the  center  is 
9i  inches,  it  has  a  force  of  12,500  pounds  acting  at  A,  37  inches  from  the  fulcrum  R. 
The  bending  moment  Mwe  have  already  found  to  be  38541.66  foot-pounds,  and 

M  =  d*  x  b  x  200. 
d2  =  9.25  x  9.25  =  85.5625. 
Now,  substituting  for  the  symbols  their  values,  we  have 

38541.66  =  85.5625  x  b  x  200 ; 
hence 

38541.66 

6  '     17112.5  =  2J 


548 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


The  lever  is  of  course  weakened  by  drilling  the  holes  for  the  fulcrum  pin,  but  this 
is  made  up  by  the  extra  thickness  allowed  in  the  center  of  the  lever  for  wear.     The 


Kg.  903. 


Fig.  903. 


Fig.  905.. 


Fig.  906. 


! -V ... 

Eig.|904 


Fig.  907 


depth  at  c  should  be  one-half  the  depth  at 
the  center. 

ENGINE-TRUCK  AXLE-BOXES. 

545.  The  construction  of  these  boxes 
varies  somewhat.  The  one  shown  in  Figs. 
902,  903,  904  is  probably  the  most  common 
one  used,  and  is  adapted  for  a  four-wheeled 
truck.  The  brass  B  is  pressed  into  the  box, 
and  the  oil-cellar  C  is  held  in  position  by 
the  two  pins  D  D.  The  depressions  E  E  receive  the  ends  of  the  equalizing  lever. 

Figs.  905,  906,  907  show  another  engine-truck  box,  which  differs  from  the  former  by 
having  the  brass  of  polygonal  form  laid  into  the  box.    Since  in  both  of  these  boxes 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


549 


the  oil-cellars  are  held  in  place  by  the  pins  D  D,  they  can  be  used  in  trucks  which  have 
no  inside  collars  on  the  axles.  These  collars  are  considered  by  many  master-mechan- 
ics to  be  of  very  little  use ;  in  fact,  for  ready  access  to  the  box,  and  easy  inspection, 
they  are  in  the  way,  and  are  therefore  not  used  on  many  axles. 

Figs.  908,  909,  910  show  a  box  in  which  the  oil-cellar  is  slid  in  end-ways  and  is 
prevented  from  slipping  out  by  the  hub  of  the  wheel  and  the  collar  on  the  axle,  which  in 
this  case  must  be  used.  The  brass  has  two  strips  of  Babbitt  metal,  F  F,  which  project 
beyond  its  bearing.  It  is  believed  that  with  Babbitt  metal  projecting  in  this  manner 
heating  of  the  journal  will  be  to  a  great  extent  prevented.  The  boxes  in  Figs.  902  and 

are  also  used  for  two- wheeled  trucks. 


ENGINE-TRUCK  AXLE  JOURNAL. 

546.  In  Table  81  we  give  the  sizes  of  engine-truck  axle  journals  as  we  have  found 
them  in  actual  service,  under  engines  built  by  different  firms.  It  will  be  noticed  that 
the  sizes  of  these  journals  vary ;  they  do  not  bear  a  certain  ratio  to  the  weight  on  the 
truck  or  size  of  engine.  Yet  this  table  will  aid  us  in  establishing  a  rule  for  finding 
t  In'  diameter  and  length  of  a  journal. 

TABLE  81. 

DIMENSIONS  OP  ENGINE-TRUCK  JOURNALS,   FOR   EIGHT-  AND  TEN-WHEELED  ENGINES,   AS  USED 

IN  ACTUAL  SERVICE. 


Size  of  Cylinders. 

Diameter  of  Journal*. 

Length  of  Journals. 

Column  1. 

Column  2. 

Column  3. 

11  x  15  inches. 

4  inches. 

6  in. 

bes. 

13  x  20 

4 

6 

14  x  20 

4 

6 

16  x  24 

4* 

• 

8 

17  x  24 

4i 

8 

18  x  24 

5 

12 

19  x  24 

5 

8 

19  x  24 

4 

81 

19  x  26 

5i 

13 

20  x  24 

5 

12 

The  surface  velocities  of  the  truck  journals  are  very  high,  varying  from  6  to  10 
feet  per  second.  For  such  high  velocities  it  is  advisable  not  to  allow  the  pressure  on 
the  journal  to  exceed  80  pounds  per  square  inch  of  projected  area.  The  pressure  on 
the  journal  is,  of  course,  found  by  subtracting  from  the  weight  on  the  truck  (as  given 
in  Tables  5  to  8)  the  weight  of  wheels  and  axles,  and  dividing  the  remainder  by  the 
number  of  journals ;  the  quotient  will  be  the  weight  on  the  projected  area  of  one 
journal,  and  dividing  this  by  80  will  give  us  the  number  of  square  inches  in  the 
projected  area.  For  many  practical  purposes  we  can  get  results  sufficiently  accurate 
by  dividing  the  whole  weight  on  the  truck,  including  the  weight  of  wheels  and  axles, 
by  the  number  of  journals,  and  then  divide  this  quotient  by  120;  the  last  quotient 
will  be  the  number  of  square  inches  in  the  projected  area  of  one  journal.  For 
example :  Referring  to  Table  5,  we  find  that  the  weight  on  truck  for  an  eight-wheeled 


550 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


passenger  engine  with  cylinders  16  x  24  inches  is  23,832  pounds ;  dividing  this  by  4, 

2383^ 
the  number  of  journals  in  the  truck,  we  have  — r —  =  5958,  which  we  may  assume  to 

be  the  pressure  in  pounds  on  the  projected  area  of  one  journal;  and  dividing  this 

5958 
quotient  by  120,  we  have  -r^r  =  49.65  square  inches  in  the  projected  area  of  one 

journal.  After  this  area  has  been  found  it  is  an  easy  matter  to  find  the  diameter, 
provided  we  know  the  ratio  of  the  diameter  and  length  of  journal.  Referring  to 
Table  81,  it  is  seen  that  a  fair  average  ratio  is  1  to  2 — that  is,  the  length  is  twice  the 
diameter.  Adopting  this  as  a  standard,  we  have  the  following  rule  for  finding  the 
diameter : 

RULE  114. — Divide  the  projected  area  of  the  journal  in  square  inches  by  2,  and 
extract  the  square  root  of  the  quotient ;  the  result  will  be  the  diameter  in  inches ;  the 
length  will  be  equal  to  twice  the  diameter. 

EXAMPLE  150. — Compute  the  diameter  and  length  of  the  engine-truck  journals  for 
an  eight-wheeled  passenger  engine  with  16  x  24  inch  cylinder. 

We  have  already  found  that  the  projected  area  of  each  journal  in  this  class  and 
size  of  engine  should  be  49.65  square  inches.  Hence,  according  to  rule,  we  have 

'9     =  24.82,  and  the  square  root  of  24.82  is  5  veiy  nearly ;  hence  the  diameter  of 

the  journal  should  be  5  inches  and  its  length  10  inches. 

In  this  way  we  have  computed  the  dimensions  of  the  journal  in  the  following 
table.  In  Column  2  the  weight  on  the  truck  is  given,  which  has  been  taken  from 
Table  5 ;  the  projected  areas  are  given  in  Column  3 ;  the  diameter  of  the  journals  in 
the  nearest  quarter  of  an  inch  in  Column  4;  and  the  length  in  Column  5. 

TABLE   82. 

DIMENSIONS   OP   ENGINE-TKUCK  JOURNALS   FOR  EIGHT-WHEELED  PASSENGER  ENGINES. 


Column  1. 

Column  2. 

Column  3. 

Column  4. 

Column  5. 

Cylinders. 

Weight  on  Truck  in 
Pounds. 

Projected  Area  of  One 
Journal  in  Square  Inches. 

Diameter  of  Journal. 

Length  of  Journal. 

Diameter. 

Stroke. 

10  inches. 

20  inches. 

10,000 

20.83 

3J  inches. 

6^  inches. 

11 

20 

13,310 

27.72 

3f 

ii 

12 

22 

14,850 

30.93 

4 

8 

13 

22 

17,070 

35.56 

4i 

Si 

14 

24 

19,242 

40.08 

4i 

9 

15 

24 

22,090 

46.02 

4* 

94 

16 

24 

23,832 

49.65 

5 

10 

17 

24 

26,010 

54.18 

5i 

10i 

18 

24 

28,680 

59.75 

5* 

11 

Comparing  these  dimensions  of  journals  with  those  in  actual  use,  as  given  in 
Table  81,  we  notice  that  they  agree  fairly  well  up  to  17  x  24  inches  cylinders.  For  the 
17  and  18  inches  cylinders  the  computed  diameters  are  somewhat  too  large  and  may 
be  reduced  £  of  an  inch.  The  diameters  of  journals  found  by  the  foregoing  computa- 
tions are  always  large  enough,  even  if  they  are  reduced  a  little,  to  resist  all  the  forces 
which  tend  to  break  the  axles ;  hence,  computations  for  strength  are  not  necessary. 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


551 


The  following  table  gives  the  dimensions  of  journals  in  two-wheeled  trucks  for 
consolidation  engines.  These  have  been  computed  in  the  same  manner  as  was 
followed  for  finding  those  of  a  four-wheeled  truck  under  an  eight-wheeled  passenger 
engine.  Of  course,  for  consolidation  engines  we  divide  the  total  weight  on  the 
journals  by  2  instead  of  4  for  finding  the  weight  on  each  journal. 

TABLE   83. 

DIMENSIONS  OP   ENGINK-TKUCK  JOURNALS   FOR  CONSOLIDATION   ENGINES. 


Column  1. 

Column  2. 

Column  3. 

Column  4. 

Column  5. 

Cylinders. 

Weight  on  Truck  in 

I'lillllil-. 

Projecied  Area  of  One 
Journal  in  Square  Inches. 

Diameter  of  Journal. 

Length  of  Journal. 

Diameter. 

Stroke. 

14  inches. 
15      " 
20      " 
22      " 

16  illrllrs. 

18      " 
24      " 
24      " 

6,272 
8,100 
14,400 
16,727 

26.13 
33.75 
60.75 
69.69 

3J  inches. 
4        " 
51      " 
5J      " 

?i  inches. 
8        " 
11         " 
Hi      " 

It  may  be  advantageous  to  show  how  the  rule  for  finding  the  diameter  and  length 
of  journals  has  been  found  when  their  ratio  is  known,  as  it  will  enable  us  to  establish 
rules  for  any  other  ratio  of  diameter  to  length.  Let  a  denote  the  number  of  square 
inches  in  the  projected  area  of  the  journal,  and  x  the  diameter  in  inches  of  the 
journal.  If,  now,  the  length  is  to  be  twice  the  diameter,  we  denote  the  length  by  2x. 

Multiplying  the  diameter  by  the  length,  we  obtain  the  projected  area.  This  con- 
dition is  expressed  in  symbols : 

x  x  2x  =  a. 

Performing  the  multiplication  as  indicated  by  the  first  member  of  the  equation, 
we  have 


from  which  we  get 


2aj*  =  a, 


and 


x  = 


This  last  expression  reads  :  The  diameter  is  equal  to  the  square  root  of  one-half 
the  area,  which  is  the  same  as  given  in  Rule  114.  If,  now,  the  length  of  journal  is  2j 
times  the  diameter,  and  using  the  same  symbols  as  before,  we  have 


x  x 


=  a 


=  a 


a 


and 


*  =  V  2V 


From  the  foregoing  we  learn  that  the  diameter  of  the  journal  is  equal  to  the 
square  root  of  the  quotient  which  is  obtained  by  dividing  the  projected  area  of  the 
journal  by  the  ratio  of  diameter  to  length. 


CHAPTER   XIV. 

OIL-CUPS.— VALVES.— COCKS— INJECTOR. 


OIL-CUPS. 

547.  A  great  diversity  of  opinion  exists  in  regard  to  the  best  form  of  oil-cups;  in 
fact,  since  there  is"  more  or  less  trouble  caused  by  the  oil-cups  breaking  off  at  their 
shanks,  they  are  sometimes  considered  to  be  a  nuisance,  and  therefore  on  a  few  roads 
oil-cups  of  any  kind  are  not  used,  the  oil  being  fed  through  simple  oil-holes.  Of  the 
various  kinds  adopted  we  shall  show  the  construction  of  the  principal  ones. 

Fig.  911  shows  the  simplest  form  of  oil-cup.  It  is  made  of  brass  in  one  piece ; 
its  reservoir  is  filled  with  waste  and  oil  This  cup  is  sometimes  used  for  link  hangers, 


Pig.  913 


rocker-boxes,  and  other  stationary  bearings.  Fig.  912  shows  an  oil-cup  suitable  for 
the  guides.  The  flow  of  oil  is  regulated  by  the  brass  spindle  A,  whose  conical  end  fits 
the  seat  in  the  bottom  of  the  cup.  For  the  purpose  of  filling  the  cup,  the  upper  part 
of  the  spindle  is  made  hollow,  from  which  a  hole  6  leads  into  the  reservoir  of  the  cup. 

Fig.  913  shows  another  form  of  guide  oil-cup.  In  this  cup  the  flow  of  oil  is  also 
regulated  by  an  adjustable  spindle  A.  A  pointer  b  is  fastened  to  the  upper  end  of 
this  spindle  for  the  purpose  of  indicating  the  extent  of  opening  for  the  oil  to  flow 
through,  and  when  set,  the  spindle  is  prevented  from  turning  by  the  springs  c  c,  which 
bear  against  the  in  side  of  the  cup.  After  the  supply  of  oil  has  been  poured  into  it 
the  whole  is  covered  by  the  brass  casing  B.  The  aim  in  the  design  of  these  cups  is, 


MODERX  LOCOMOTIVE   CONSTRUCTION. 


553 


of  course,  to  feed  the  oil  gradually,  and  keep  the  slides  constantly  and  regularly 
lubricated. 

Fig.  914  shows  the  Ricker  oil-cup.  It  is  used  on  the  rear  ends  of  the  main- 
n«ls  itud  on  the.  ends  of  the  side-rods.  The  oil  is  thrown  upwards  to  the  top  of 'the 
tube  a  by  the  motion  of  the  rods,  and  its  flow  is  regulated  by  the  needle  on  the  end  of 
spindle  A. 

Fig.  915  shows  a  cup  after  the  pattern  of  the  Nathan  Manufacturing  Co.'s  patented 
oil-oup.  A  glass  shell  B  is  placed  inside  of  the  brass  casing  A,  and  the  whole  is  made  oil- 


- 
crrj 


Fijj.  916. 


Fig.  914 


tight  by  cork  packing  placed  at  the  top  and  bottom  of  the  glass.  When  the  cup  is  to 
be  filled  the  nut  C  must  be  taken  off.  A  small  air-hole  should  be  drilled  through  the 
cover  D  to  admit  air  on  top  of  the  oil.  When  these  cups  are  designed  for  the  rear  end 


Fig.  917 


of  the  main-rods  and  for  the  side-rods,  they  have  a  steel  spindle  rf,  which  works  up 
and  down,  its  movement  being  due  to  the  motion  of  the  rods,  thereby  feeding  a  certain 
amount  of  oil  regularly  onto  the  crank-pin  journals.  The  lift  of  these  spindles,  and 
consequently  the  amount  of  oil  to  be  fed,  is  regulated  by  the  screw  e.  When  the  cup 


554 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


is  at  any  time  taken  off  the  rods,  the  spindle  d  is  prevented  from  falling  out  of  the  cup 
by  a  small  pin  driven  through  its  upper  end.  The  advantage  of  these  cups  is  that  the 
oil  in  the  reservoir  is  constantly  exposed  to  view. 

Since  much  trouble  is  experienced  with  the  oil-cups  breaking  off  at  the  shanks, 
some  master-mechanics  fasten  the  cups  in  the  manner  shown  in  Fig.  916.  On  some 
roads  the  rod  oil-cups  are  forged  to  the  rods  or  straps,  as  shown  in  Tig.  917. 

An  excellent  way  of  keeping  the  crank  pins  and  side-rod  pins  lubricated  is  by  the 
use  of  an  oil-cellar  fastened  to  the  bottom  of  the  strap,  as  shown  in  Fig.  918.  This 
cellar  is  filled  with  waste  and  oil,  the  waste  extending  through  the  oil-hole  up  to  and 
touching  the  crank  pin ;  the  capillary  force  draws  the  oil  up  the  waste  and  lubricates 
the  pins. 

CYLINDER  OIL-CUPS. 

548.  Figs.  919,  920  show  an  outside  view  and  section  of  an  ordinary  cylinder  oil- 
cup  ;  one  of  these  is  used  for  each  cylinder.  The  shank  D  is  usually  attached  to  the 

side  of  the  boiler  inside  of  the  cab  by  nipples  or 
elbows,  so  that  the  cup  stands  in  a  vertical  position. 
A  copper  oil-pipe  F,  usually  -£$  or  x«  inch  outside 
diameter,  is  connected  to  the  shank  E ;  the  other 
end  of  this  pipe  is  connected  to  the  steam-chest  plug 
shown  in  Fig.  923,  and  described  further  on. 
The  conical  end  of  the  spindle  B  (Fig.  920) 
acts  as  a  valve  by  which  the  oil-hole  H  in 
the  stem  C  can  be  opened  or  closed.  The  end 
of  the  stem  C  has  attached  to  it  a  valve  C2 
for  admitting  steam  into  the  oil-pipe  F  for 
the  purpose  of  blowing  out  any  matter 
that  may  collect  in  it  from  time  to  time. 
Of  course  the  valve  C2  is  always  closed 
when  oil  is  to  be  fed  into  the  cylinder.  The 
oil  is  poured  into  the  cup  A,  and  when 
the  end  of  the  spindle  B  is  lifted  off  the 
seat  on  //  the  oil  can  flow  into  the  cham- 
ber G  and  into  the  oil-pipe  F;   but  it  can- 
not flow  into  the  steam-chest  and  cylinder 
1 1 1»  until  the  throttle  valve  is  closed,  because 

®J  Fig.  919  |4J     Fig.  920  go  iong  as  there  is  a  steam  pressure  in  the 

steam-chest  the  valve  in  the  steam-chest 

plug  (Fig.  923)  will  be  kept  closed ;  but  as  soon  as  the  throttle  valve  is  closed  a  partial 
vacuum  will  be  formed  in  the  steam-chest,  and  while  the  engine  is  still  in  motion  the 
oil  in  the  pipe  F  will  be  sucked  in  the  steam-chest  and  cylinder.  Whenever  the  valve 
C2  is  to  be  opened  care  must  be  taken  to  first  close  the  valve  on  the  hole  H,  other- 
wise the  engineer  is  liable  to  be  scalded. 

Figs.  921,  922  show  another  cylinder  oil-cup  somewhat  different  in  design  from  the 
foregoing,  but  both  work  on  the  same  principle.  The  shank  D  is  attached  to  the  boiler, 


MODERN  LOCOMOTIVE   COSSTRVCTWN. 


555 


and  the  oil-pipe  is  connected  to  the  shank  E.  The  cup  A  is  cast  to  the  stern  B,  whose 
bottom  end  acts  as  a  valve  and  closes  or  opens  the  hole  leading  into  the  chamber  G. 
The  cup  A  also  forms  a  handle  for  turning  the  stem  B.  Oil  is  poured  into  the  cup  A, 
from  whence  it  flows  into  the  cup 
below,  and  when  the  valve  B  is  opened 
the  oil  is  drawn  into  the  steam-chest 
as  soon  as  the  throttle  valve  is  closed. 
A  valve  is  fitted  into  the  end  of 
the  stem  C  and  admits  steam  into 
the  oil-pipe  for  the  purpose  of  clean- 
ing it. 

Fig.  923  shows  a  steam-chest 
plug;  it  is  usually  made  of  brass. 
Its  shank  C  is  screwed  into  the  steam- 


chest  cover,  and  the  nut  D  holds 
down  the  steam-chest  false  cover. 
The  oil-pipe  is  connected  to  the  up- 
per end  of  the  plug  by  the  nut  A. 
This  end  contains  a  small  valve  B, 
which  is  free  to  move  up  or  down. 
When  the  engine  is  working,  the 
steam  in  the  chest  lifts  the  valve  B, 
presses  it  against  the  end  of  the  oil- 
pipe  and  prevents  the  steam  from  en- 
tering it.  From  the  foregoing  it  will 
be  seen  that  the  cylinder  can  be  lu- 
bricated only  when  the  steam  is  shut 
off,  as  we  have  stated  in  the  descrip- 
tion of  the  cylinder  oil-cups,  and  then 
nearly  the  whole  quantity  of  oil  in 
the  cup  will  be  drawn  into  the  cylin- 
der during  a  few  revolutions  of  the 
wheels,  making  the  feed  irregular  and 
not  nearly  as  constant  as  is  desirable. 
The  small  valve  B  in  the  plug  not 
only  relieves  the  oil-pipes  of  consid- 
erable internal  pressure,  but  it  also 
serves  another  useful  purpose.  If  the 
small  valve  B  is  not  used,  the  pipes 
will  be  continually  full  of  steam, 
which  is  liable  to  condense  and  be 
forced  out  with  the  oil  through  the 
cups  in  case  these  are  opened  before 
the  pressure  in  the  steam-chest  has  been  reduced, 
is  prevented  by  the  use  of  the  small  valve  B. 


3 


L "_  _'jf ~ =: 

„  i~HEfek  9 


The  occurrence  of  this  nuisance 


556 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


SIGHT-FEED   LUBRICATORS. 

549.  During  recent  years  sight-feed  lubricators  for  the  cylinders  have  to  a  great 
extent  displaced  the  ordinary  cylinder  oil-cups  previously  described.  Several  kinds 
are  manufactured. 

Figs.  924,  925,  926,  and  926A  show  one  of  these  made  and  patented  by  the  Nathan 
Manufacturing  Company,  of  New  York.  With  these  lubricators  the  feed  is  regular  and 


Fig.  924 


Pig.  925 


Fig.  "926 


C  C 

Fig.  926A 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


557 


coiitiniious,  whether  steaming  or  with  steam  shut  off,  and  when  going  up  or  down 
grade.  The  flow  of  oil  is  constantly  in  sight,  showing  at  all  times  whether  the  lubri- 
cator is  working  satisfactorily  or  not.  The  lubricators  are  placed  inside  of  the  cab,  and 
are  usually  attached  to  the  rear  end  of  the  boiler,  as  shown  in  Fig.  926,  within  sight 
and  easy  reach  of  the  engineer  or  the  fireman.  The  principle  upon  which  it  works  is 
that  the  water,  which  is  heavier  than  the  oil,  displaces  the  oil  in  the  cup,  causing  it  to 
flow  drop  by  drop  through  a  body  of  water  in  the  sight-feed  glass,  and  then  it  enters 
the  oil-pipes  which  lead  to  the  steam-chest.  Steam  is  taken  from  the  top  of  boiler  or 
from  the  dome,  and  is  conducted  into  the  condenser  JK,  where  it  is  condensed  and 
furnishes  the  working  water  column.  The  water  is  led  to  the  bottom  of  the  reservoir  / 
by  aii  inside  pipe ;  its  flow  is  regulated  by  the  valve  Z>,  Fig.  925.  The  oil  which  floats 
on  the  water  in  the  reservoir  /  is  carried  upwards  until  it  enters  the  top  of  the  pipe  P, 
from  whence  it  flows  downwards,  then  through  the  channel  -7,  and  finally  enters  into 
the  bottom  of  the  sight-feed  glass  K,  through  which  it  flows  upwards  drop  by  drop, 
and  then  enters  the  oil-pipe  which  leads  the  oil  into  the  steam-chest  and  cylinders. 
The  quantity  of  oil  entering  the  sight-feed  glass  is  regulated  by  the  valves  C,  Fig.  924. 
Inside  of  the  condenser  E  there  are  two  entirely  separate  steam  conduits  L,  which 
allow  the  minimum  quantity  of  steam  to  enter  into  oil-pipe  H  (see  Fig.  926) ;  this 
steam  becomes  saturated  with  the  oil  and  forms  a  steam  lubricant.  These  lubrica- 
tors are  arranged  to  form  two  distinct  oilers,  one 
for  each  cylinder,  thereby  avoiding  the  possibility 
of  feeding  all  the  oil  into  one  cylinder ;  each  cylin- 
der is  lubricated  independently  of  the  other.  Each 
side  is  also  furnished  with  an  independent  hand  or 
auxiliary  oiler  0  0;  these  work  on  the  same  prin- 
ciple as  the  ordinary  cylinder  oil-cups  shown  in 
Fig.  921 ;  they  communicate  directly  with  the  out- 
let passages  of  the  lubricator,  and  are  used  in  case 
a  sight-feed  or  gauge  glass  has  been  broken.  Such 
an  accident  necessitates  the  shutting  off  of  the 
sight  lubricator.  The  glass  tube  G  shows  the  height 
of  water  in  the  reservoir  7.  The  cock  Wis  simply 
a  waste  cock  for  draining  the  reservoir  when  nec- 
essary. A  is  the  filling  plug,  and  />',  Fig.  926,  the 
steam  valve.  The  safety  valves  F  F  are  always 
kept  open,  except  when  one  of  the  glasses  is  broken. 
In  such  cases  the  valves  /•',  />,  and  II  must  be  closed, 
so  as  to  .shut  off  the  lubricator  and  allow  the  cylin- 
ders to  be  lubricated  by  the  auxiliary  oilers  0  0. 

The  oil-pipe  77,  one  on  each  side  of  the  engine, 
is  run  along  the  boiler  underneath  the  lagging,  and 
is  connected  to  the  steam-chest  plug.  For  sight 

lubricators  the  valve  7>  in  the  steam-chest  plug  (Fig.  92o)  must  be  removed  so  as  to 
maintain  the  proper  lubrication  when  the  engine  is  running. 

Fig.  926B  shows  a  combined  sight-feed  lubricator  for  oiling  the  cylinders  and  air- 


558 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


brakes,  and  is  a  simple  and  effective  device  for  the  purpose.  Its  general  appearance 
and  construction  does  not  differ  much  from  that  of  the  lubricator  shown  in  Fig.  924, 
and  works  on  the  same  principle.  The  same  letters  in  all  these  lubricators  indicate 
similar  parts. 


BLOWER  VALVES. 


550.  Fig.  927  shows  the  construction  of  an  ordinary  blower  valve.     Its  shank  A  is 
either  screwed  into  the  boiler  shell  or  into  the  steam  chamber  shown  in  Fig.  929.     The 


JSTh'.N 


Pig.  928  A 


blower  pipe  is  attached  to  the  nipple  B.  This  pipe  is  usually  made  of  iron,  sometimes 
of  brass,  f  inch  inside  diameter.  It  is  placed  along  the  side  of  the  boiler  underneath 
the  lagging,  and  terminates  in  the  smoke-box.  Its  duty  is,  of  course,  to  create  a  draft 
when  the  engine  is  not  running. 

INJECTOR. 

551.  Up  to  1853,  when  the  Giffard  injector  made  its  appearance,  pumps  were  used 
exclusively  for  feeding  locomotive  boilers.  In  1860  this  injector  was  introduced  by 
Messrs.  Wm.  Sellers  &  Co.,  of  Philadelphia,  Pa.,  who  had  also  added  many  improve- 
ments to  the  original  instrument.  Its  advantages  were  recognized,  and  it  gained  the 
confidence  of  practical  men,  gradually  increasing  in  favor,  and  finally  superseded  the 


\IOIH:I;N  LOCOMOTIVE  CONSTRUCTION.  559 

pumps,  so  that  now  wo  rarely  find  a  locomotive  which  is  not  equipped  with  injectors, 
generally  one  on  each  side  of  boiler,  in  place  of  pumps. 

Practical  tests  and  the  results  of  many  years  of  constant  service  under  varying  and 
trying  conditions  have  placed  the  injectors,  with  their  various  modifications  as  now 
made  by  different  manufacturers,  beyond  the  experimental  stage ;  and  the  old  notions 
that  the  operation  of  an  injector  is  incomprehensible,  that  its  parts  are  liable  to  be 
impaired  and  involve  nice  adjustment  to  meet  the  practical  requirements,  have  been 
removed  by  familiarity  with  the  instrument.  Its  action  is  reliable,  its  construction 
is  simple  and  compact,  it  can  be  easily  examined,  the  cost  of  repairs  is  comparatively 
low,  and  the  amount  of  feed  water  to  be  fed  into  the  boiler  can  be  regulated  with  it  as 
well  as  with  an  ordinary  pump.  Its  inherent  advantages  are  that  the  feed  water  enters 
the  boiler  comparatively  hot  without  the  use  of  a  feed- water  heater,  thereby  reducing 
the  liability  of  subjecting  the  boiler  to  an  undue  sti'ess  as  might  occur  by  feeding 
colder  water  into  the  boiler ;  it  is  an  independent  feeder,  as  with  it  a  certain  quantity 
of  water  can  always  be  supplied  to  the  boiler  when  the  speed  of  the  engine  is  irregular, 
or  even  when  the  engine  is  standing  still ;  it  requires  no  mechanism,  such  as  eccentrics 
or  cranks,  or  other  attachments,  for  its  operation ;  and  it  is  easily  applied  to  a  locomo- 
tive without  requiring  any  change  in  the  mechanism  of  the  engine. 

The  numerous  kinds  of  injectors  now  made  are  the  developments  of  the  element- 
ary form  indicated  in  Fig.  928A,  in  which  all  the  minor  details  have  been  omitted 
so  as  to  show  plainly  the  principles  of  its  construction  and  action.  It  consists  of 
a  receiving  tube  R,  a  combining-  tube  (7,  and  a  delivery  tube  D.  These  parts  are 
common  to  all  forms  of  injectors,  with  various  modifications.  The  steam  pipe  from 
the  boiler  leading  the  steam  to  the  injector  is  attached  to  the  end  £;  the  suction  pipe 
supplying  water  to  the  instrument  is  attached  to  the  nozzle  W;  and  the  delivery  pipe 
leading  to  the  boiler  is  attached  to  the  end  13.  The  action  of  the  injector  may  be 
briefly  described  as  follows :  Steam  from  the  boiler  enters  at  S  and  flows  into  the 
receiving  tube  R  with  a  great  velocity,  causing  a  partial  vacuum  in  the  chamber  A ; 
this  causes  the  water  to  flow  through  the  nozzle  W  and  around  the  receiving  tube, 
from  whence  it  is  drawn  into  the  combining  tube  C,  where  it  combines  with  the  steam, 
and  then  flows  through  the  delivery  tube  into  the  delivery  pipe  and  is  led  to  the  check 
valve,  which  is  usually  fastened  to  the  side  of  the  boiler,  and  forces  its  way  into  the 
boiler.  If  the  supply  of  water  is  too  great,  a  waste  will  occur  at  the  overflow  0]  and 
if,  on  the  other  hand,  the  supply  of  steam  is  too  great,  air  will  be  drawn  through  the 
same  opening,  which  may  cause  a  break  in  the  feed. 

At  a  casual  glance  it  may  appear  to  be  impossible  for  an  issuing  jet  of  steam  to 
force  its  way  back  into  its  own  boiler,  but  this  nevertheless  does  occur,  and  this  action 
may  be  explained  as  follows : 

We  may  assume  that  the  steam  pipe  leading  to  the  injector  and  the  delivery  pipe 
leading  from  the  injector  to  the  boiler  form  one  continuous  pipe.  Here,  then,  we  have 
a  pipe  into  which  steam  enters  at  one  end  and  water  tends  to  enter  at  the  other  end,  and 
both  the  steam  and  water  are  subjected  to  the  same  pressure.  Under  these  conditions 
the  velocity  of  the  steam  is  much  greater  than  that  which  the  water  would  have  it'  it 
could  flow  into  the  pipe.  If,  now,  the  steam  during  its  flow  is  condensed,  and  the 
velocity  of  the  water  due  to  the  condensation  is  not  reduced  to  the  velocity  which  the 


560 


MODERN  LOCOMOTIVE    CONSTKVCTIOX. 


Cylr.oll  cup 


water  at  the  opposite  end  of  the  pipe  would  have,  then  the  water  due  to  the  condensa- 
tion will  have  the  greater  momentum  and  will  be  enabled  to  overcome  the  force  of 
water  which  tends  to  flow  from  the  boiler,  and  the  water  due  to  condensation  will  force 
its  way  into  the  boiler.  This  is  exactly  what  occurs  when  an  injector  is  interposed ; 
the  steam  enters  the  injector  at  a  very  high  velocity,  is  condensed  in  the  injector  with- 
out losing  much  of  its  velocity,  some 
of  which  is  imparted  to  the  water  which 
rushes  into  the  injector  through  the 
nozzle  TF;  the  resulting  velocity  is  of 
course  less  than  that  of  issuing  jet  of 
steam,  but  it  is  still  greater,  and  con- 
sequently its  momentum  is  also  greater 
than  the  velocity  of  the  water  tending 
to  flow  out  of  the  boiler  would  be.  This 
enables  the  steam,  with  an  additional 
amount  of  water,  to  overcome  all  re- 
sistances and  force  its  way  back  into 
its  own  boiler. 

There  are  many  different  forms  of 
injectors  manufactured,  which  need  not 
be  described  here,  as  full  information 
in  regard  to  them,  with  illustrations, 
can  be  obtained  in  the  many  excellent 
catalogues  distributed  by  the  manufact- 
urers. 

With  the  higher  steam  pressures, 
say  150  pounds,  the  temperature  of  the 
feed  water  delivered  by  the  injector  may 
reach  120  degrees  Fahrenheit,  and  for  very  low  pressures  it  may  reach  130  degrees. 

When  the  injector  is  used  as  a  boiler  feeder  it  is  a  very  efficient  instrument,  and 
it  will  require  less  of  the  fuel  than  a  pump  of  the  same  capacity.  The  heat  imparted 
to  the  feed  water  is  not  lost,  as  it 
is  all  returned  to  the  boiler.  But 
if  an  injector  is  used  simply  as  a 
pump — that  is  to  say,  for  raising 
water  or  similar  purposes — the 
heat  imparted  to  the  delivery 
water  is  wasted,  and  on  this  ac- 
count the  efficiency  of  the  injector 
is  lower  than  that  of  a  pump ;  or, 
in  other  words,  it  requires  more 
steam  to  raise  a  given  amount  of  water  than  will  be  required  for  running  a  pump 
doing  the  same  amount  of  work. 

Steam  for  working  the  injector  should  always  be  taken  from  the  dome  or  the  highest 
part  of  the  boiler  so  as  to  obtain  dry  steam,  as  wet  steam  is  liable  to  injure  it. 


Fig.1930 


.U'</)/./,'.V  LOCOMOTIVE  CONSTRUCTION. 


561 


A  stop  valve  should  always  be  placed  between  the  steam  space  in  the  boiler  and  the 
injector.  Fig.  928  shows  the  construction  of  such  a  valve.  The  nipple  B  is  either 
attached  to  the  boiler  or  to  the  steam  chamber,  and  the  pipe  leading  to  the  injector  is 
jittarhfd  to  the  nipple  A.  The  size  of  this  valve  depends  on  the  size  and  kind  of 
injector  used.  All  pipes,  whether  steam,  water  supply,  or  delivery,  must  be  of  the 
san  10  internal  diameter  as  the  hole  in  the  corresponding  branch  of  the  injector,  and  all 
pipes  should  be  as  straight 
as  practicable.  When  steam 
is  taken  from  the  dome  a 
dry  pipe  is  generally  re- 
quired to  lead  the  steam  to 
the  stop  valve. 

For  non-lifting  inject- 
ors a  small  cock  in  the 
suction  pipe  is  required 
for  regulating  the  supply 
of  water.  This  cock  gen- 
erally takes  the  place  of 
the  feed  cock  shown  in 
Fig.  578,  and  is  worked  by 
a  suitable  mechanism  from 
the  foot-board. 

552.  We  have  already 
shown  one  form  of  steam 
chamber  in  Fig.  536.  All 
the  valves  required  in  the 
cab  are  screwed  into  this 
chamber.  Fig.  929  shows 
another  form  of  a  steam 
chamber;  it  is  generally 
used  on  boilers  which  do 
not  extend  very  far  into  the 
cab.  The  pipe  A  is  usually 
led  into  the  dome  so  that 
dry  steam  will  be  supplied 
to  all  the  valves  attached 
to  this  chamber.  The  sup- 
ply of  steam  can  be  shut 
off  by  the  valve  B,  for  the  purpose  of  repairing  any  one  of  the  valves  attached  to  the 
chamber  when  steam  is  in  the  boiler.  The  flange  C  is  bolted  to  the  boiler  by  means  of 
studs,  and  the  joint  between  the  shell  and  chamber  is  a  ball  joint,  as  shown.  Tho  dotted 
flange  shows  the  form  when  the  steam  chamber  has  to  bo  bolted  on  the  side  of  the  boiler. 

•">3.  Fig.  9.30  shows  the  construction  of  an  ordinary  gauge  cock ;  it  is  made  of 
brass,  and  usually  three  cocks  are  used  for  each  boiler.  They  are  frequently  screwed 
into  the  rear  head  of  the  boiler,  but  sometimes  into  the  side.  The  center  of  the  lowest 


K— 3*- 
Fig.  933 


562  MODERN  LOCOMOTIVE    CONSTRUCTION. 

gauge  cock  should  be  about  2  inches  above  the  highest  point  of  the  crown  sheet,  and 
the  upper  cock  should  be  placed  a  little  above  the  intended  water  level.  This  will 
usually  bring  the  gauge  cocks  from  3  to  4  inches  apart.  Consequently,  when  the 
upper  gauge  cock  discharges  water  the  indications  are  that  there  is  too  much  water  in 
the  boiler ;  and  it  scarcely  need  be  said  that  there  is  danger  when  the  lower  cock 
discharges  steam. 

In  trying  the  gauge  cocks  the  water  is  discharged  into  a  drip  pan,  from  whence  it 
is  led  through  a  pipe  to  the  road  bed. 

554.  Figs.  931,  932,  933,  934  show  different  views  of  a  water-gauge  glass ;  the  lower 
cock  B  is  in  communication  with  the  water  in  the  boiler,  and  the  upper  cock  C  com- 
municates with  the  steam.  The  glass  tube  A  is  usually  from  11  to  15  inches  long,  its 
outer  diameter  is  generally  |  inch,  and  its  thickness  £  inch.  Steam-  and  water-tight 
joints  around  the  glass  are  secured  by  rubber  packing.  The  lower  cock  B  should  be 
placed  so  that  the  lower  end  of  the  glass  at  a  will  indicate  about  2  inches  of  water 
above  the  highest  part  of  the  crown  sheet.  The  cock  D  is  used  for  blowing  out  the 
sediment  which  is  liable  to  collect  in  the  glass.  The  balls  d  d  are  for  the  purpose 
of  shutting  off  automatically  the  steam  and  water  in  case  the  glass  has  been  broken. 
When  everything  is  in  working  order  the  balls  d  d  roll  away  from  their  seats  and 
allow  a  free  passage  for  the  steam  and  water  into  the  glass.  But  should  the  glass 
break,  the  swift  currents  of  the  steam  and  water  carry  the  balls  against  the  seats, 
thereby  preventing  any  further  escape.  When  a  new  glass  has  been  put  in  the  balls 
are  forced  off  the  seats  by  the  valve  stems. 


CHAPTER  XV. 

TENDERS.— TENDER  TRUCKS. 

TENDERS. 

555.  Fig.  935  shows  a  side  elevation  of  a  tender ;  its  frame  is  made  of  wood.     All 
tender  frames,  whether  made  of  wood  or  iron,  are  covered  with  planks,  generally  2 
inches  thick ;  they  form  a  floor  on  which  the  tank  is  placed.     The  usual  method  of 
fastening  the  tank  to  the  frame  is  to  rivet  a  wrought-iron  bracket  or  knee  a  a  to  each 
outer  corner  of  the  tank,  and  three  or  four  brackets  of  the  same  kind  are  riveted  to 
tlic  sheets  which  form  the  coal  space;  a  bolt  through  the  frame  and  foot  of  each 
bracket  holds  the  tank  securely  in  place.     Care  should  always  be  taken  to  secure  the 
tank  in  such  a  way  as  will  not  cause  it  to  be  subjected  to  a  longitudinal  stress  due  to 
the  pull  of  the  engine.     If  the  tank  has  to  resist  part  of  the  pull  of  the  engine,  leakage 
is  liable  to  occur. 

When  a  wooden  frame  is  used  the  front  and  back  draw-heads  are  bolted  to  it  by 
four  bolts  b  b,  which  extend  through  the  whole  length  of  frame  and  must  be  made 
strong  enough  to  transmit  the  whole  pull  of  the  engine  to  the  cars  behind  it.  The 
members  of  a  wooden  frame  should  never  be  subjected  to  a  tensile  stress ;  the  only 
stress  in  a  wooden  tender  frame  should  be  that  due  to  the  weight  of  the  tank,  water 
and  fuel,  and  the  push  of  the  engine. 

556.  The  diameters  of  the  longitudinal  bolts  b  b  are  easily  found.    The  tension  in 
these  bolts  cannot  be  greater  than  that  due  to  the  adhesion  of  the  driving  wheels.    If, 
for  instance,  the  weight  on  the  drivers  is  60,000  pounds,  and  if  we  assume  the 
adhesion  to  be  equal  to  J  of  the  weight  on  the  drivers,  then  the  tension  in  the  four 

60000  15000 

bolts  will  be      -7 —  =  15000  pounds,  and  the  tension  in  one  bolt  will  be   —7 —  =  3750 

pounds.  This  tension  or  pull  is  frequently  applied  suddenly,  and  therefore  the  normal 
stress  per  square  inch  of  cross-section  of  these  bolts  should  not  exceed  3,000  pounds, 

hence  the  smallest  cross-sectional  area  of  each  bolt  should  be  "o^j™  =  1.250  square 

inches ;  the  corresponding  diameter  is  1 J  inches ;  the  smallest  cross-section  is  at  the 
bottom  of  the  thread,  hence  this  section  should  be  1J  inches  diameter,  which  will  give 
us  a  little  over  lj|  inches  for  the  diameter  outside  of  thread. 

557.  The  iron  trucks  used  under  this  tender  (Fig.  935)  are  alike  in  every  respect, 
excepting  that  the  rear  truck  has  two  side  bearings  r  >',  on  which  the  tender  frame 
rests;  the  front  truck   lias  no  side  bearings,  the  front  end  of  tender  frame  being 
supported  by  the  center  pin  s,  so  that  virtually  the  frame  and  tank  with  water  and 


564 


MODERX  LOCOMOTIVE   CONSTRUCTION. 


fuel  are  supported  by  three 
points  only.  Although  this 
is  good  practice,  it  is  by  no 
means  a  universal  one,  for  un- 
der many  tenders  both  trucks 
have  side  bearings  and  con- 
sequently support  the  frame 
and  tank  by  four  points.  The 
objection  to  four  points  of 
support  is  that  when  the  ten- 
der is  running  over  any  un- 
evenness  of  the  track  some  of 
the  truck  springs  are  liable 
to  be  compressed  to  a  hurtful 
extent  and  the  tender  frame 
is  liable  to  be  twisted.  With 
three  supports  the  load  will 
be  more  evenly  distributed 
over  all  the  springs  under  any 
conditions. 

Trucks  which  have  side 
bearings  like  the  rear  truck  in 
Fig.  935  are  called  side-bearing 
trucks ;  and  trucks  whose  cen- 
ter pin  acts  as  a  support,  like 
the  front  truck  in  Fig.  935,  are 
called  center-bearing  trucks. 

558.  Safety  chains  t  t  are 
bolted   to  the    truck   frames 
and  the  sides  of  tender  frame. 
The  wooden  brake  beams  u  u2, 
etc.,  are  supported    by  their 
hangers  o  o;   they  also  have 
safety  chains  p  p  which  can 
support  the  beams  in  case  the 
hangers  break. 

559.  The  brakes  in  Fig. 
935  are  designed  to  be  worked 
by  hand.  One  end  of  the- chain 
v  is  attached  to  the  bottom  of 
the  brake  shaft  c,  the  other 
end  of  the  chain  is  connected 
to  the  rod  h.     In  turning  the 
shaft  c  by  the  hand-wheel  d 
the  chain  v  is  wound  around 


Moi>i:i;.\ 


565 


Fig.  939 


Fig.  938    / 


the  bottom  of  the  shaft,  thereby  pulling  the  rod  /*  towards  the  front  of  tender.  This 
rod  is  connected  to  the  rear-brake  lever  m  on  the  front  truck ;  the  rod  q  connects 
the  lower  end  of  the  lever  m  and  the  lower  end  of  its  mate  n ;  the  fulcrums  of  these 
levers  are  supported  by  the  wrought-iron  jaws  through  the  center  of  the  brake 
beams.  The  upper  end  of  the  lever  n  is  connected  to  the  rod  <t,,  which  is  in  turn 
connected  to  the  lever  k  on  the  rear  truck.  The  lower  end  of  k  is  connected  to  the 
lower  end  of  the  lever  I  by  the  rod  j,  and  the  upper  end  of  I  is  held  by  the  rod  i, 

which  is  fastened  to  the  tender  frame.    Now,  it  will  readily 

be  seen  that  a  pull  on  the  rod  h  will  put  all  the  brakes  in 

action ;  those  on  the  front  truck  act  before  those  on  the  rear 

truck. 

The  brake  shaft  c  is  shown  in  detail  in  Fig.  936.     Its 

upper  end  is  held  in  position  by  a  cast-iron  bracket  bolted  to 

the  tank ;  separate  views  of  this  bracket  are  shown  in  Fig. 

937.     The  lower  end  of  the  shaft  c  is  supported  by  a  step 

bearing,  which  is  usually  made 

of  wrought-iron;  it  is  bolted 

to  the  bottom  of  the  tender 

frame,  and  the  shaft  is  pre- 
vented from  moving  upwards 

by  a  key  shown  in  Fig.  938. 

The  hole  d  lakes  the  eye  bolt 

shown  in  Fig.  939,  to  which 

the  chain  is  attached. 

The  casting  shown  in  Fig. 

940   is  bolted  to  the  top  of 
936     tender  floor.    The  brake  shaft 

passes    through   the   hole   a. 

Just    above    this    hole,    the 

ratchet  wheel  which  is  shown 

in  Fig.  941  is  keyed   to  the 

brake  shaft;  this  wheel  en- 
gages with  the  pawl  shown  in  Fig.  942 ;  it  works  on  the  pin  b  (Fig.  940).  The  object 
of  the  pawl  and  ratchet  wheel  is,  of  course,  to  prevent  the  brake  shaft  from  turning 
backwards,  thereby  holding  the  brakes  in  action.  The  pawl  is  worked  by  foot. 

560.  Figs.  943  to  946   show  separate  views  of  the  wooden   tender  frame.     Its 
construction  is  so  plainly  shown  that  further  description  is  unnecessary. 

561.  The  draw-head  at  the  rear  end  of  the  tender  frame  has  two  pockets,  as 
indicated  in  Figs.  944,  945,  by  which  the  tender  can  be  coupled  to  cars  of  different 
heights.   The  number  of  pockets  in  the  rear  draw-head  depends  much  on  the  service  for 
which  the  engine  is  designed.     In  many  engines  this  draw-head  has  only  one  pocket, 
and  in  switching  engines  three  or  even  a  greater  number  of  pockets  may  be  required. 

The  hole  in  this  head  for  llie  coupling  pin  is  a  cored  hole,  and  the  pin  is  smooth 
forged;  the  pin  has  always  a  little  play  in  the  hole — in  fact,  more  play  than  is  admis- 
sible in  the  front  draw-head. 


Fig.  941 


Fig.  942 


u 


Fig.  010 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


567 


Figs.  947  to  950  show  different  views  of  the  front  draw-head.  The  designs  of 
these  heads  differ  in  the  various  classes  of  engines.  The  one  here  shown  was  adopted 
for  the  tender  illustrated  in  Fig.  935.  The  hole  for  the  coupling  pin  is  bored,  and  the 
pin  is  turned  so  as  to  reduce  the  play  between  engine  and  tender  as  much  as  possible. 


TANK. 


562.  Figs.  951  to  953  show  separate  views  of  the  tank  and  the  manner  of  bracing 
its  sides  and  ends.     Some  of  the  braces  are  made  to  serve  as  swash  plates,  to  prevent 


Fig.  956 

f k-x- 


Fig.947 

, q«4- — 


Fig.  948 


Fig.  954. 


Fig.  950 


Fig.  957. 


j 

Fig.  958    \J 


a  violent  motion  of  water  when  the  engine  is  suddenly  stopped  or  started.  The  tank 
is  filled  through  the  manhole  C,  whose  cover  is  made  of  plate-iron,  with  a  wooden 
block  riveted  to  its  under  side  as  shown.  The  wings  D  are  riveted  to  the  sides  and 
rear  end  of  tank,  chiefly  for  the  purpose  of  preventing  fuel  and  water  from  dropping 
off  the  tank  and  destroying  the  outside  appearance.  The  angle  irons  w  w  riveted  to 
the  legs  of  the  tank  form  a  channel-way  in  which  boards  extending  across  the  coal 
space  are  placed.  The  tank  is  generally  made  of  iron,  sometimes  of  steel.  The  outer 
sides  are  usually  -^  inch  thick,  plates  around  the  fuel  space  J  inch  thick,  and  the  top 
and  bottom  sheets  are  also  J  inch  thick.  The  water  capacities  of  tanks  vary  from 
1,500  to  3,600  gallons,  to  suit  the  size  of  engine,  and  the  fuel  space  will  hold  from  2  to 
5  tons  of  coal. 

The  hole  y  through  the  bottom  sheet  of  tank  (Fig.  951)  is  the  outlet  for  the  feed 
water ;  the  hole  x  in  the  top  of  tank  allows  the  tank  valve-rod  to  pass  through  the 
sheet ;  the  centers  of  the  holes  x  and  y  lie  in  the  same  vertical  line. 

563.  Fig.  954  shows  the  tank  valve  seat ;  its  hollow  hub  a  forms  a  guide  for  the 
valve  b  shown  in  Fig.  955.  This  seat  and  valve  is  covered  with  a  cast-iron  strainer 
shown  in  Fig.  956;  its  purpose  is  to  prevent  any  obstruction  from  entering  the  feed 
pipe  which  may  render  the  pump  or  injector  useless.  The  cast-ii-on  nozzle  shown  in 


i  I     VTfe       > 

._^S^ 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


569 


Fig.  957  is  bolted  to  the  under  side  of  tank  in  line  with  the  valve  seat.  By  means 
of  a  horned  nut  (Fig.  958)  a  cast-iron  or  brass  nipple  is  attached  to  the  nozzle.  The 
general  appearance  of  the  whole  arrangement  is  shown  in  Fig.  959.  A  rubber  hose 
is  fastened  to  the  nipple  and  leads  the  feed  water  to  the  feed  pipe  on  the  engine. 

Fig.  960  shows  the  tank  valve-rod ;  the  jaw  at  its  lower  end  is  connected  to  the 
tank  valve.     This  rod  passes  through  the  cast-iron  stand  A  shown  in  Fig.  961.     This 


Fig.  980. 


"Eg.  959.. 


stand  is  bolted  to  the  top  of  tank ;  its  upper  end  is  bored  out  larger  than  the  body, 
and  is  threaded  for  the  nut  or  gland  B,  which  presses  the  packing  against  the  valve- 
rod  and  makes  a  water-tight  joint. 

A  cast-iron  handle  C  is  fitted  and  fastened  to  the  square  end  of  the  valve-rod. 
The  bottom  of  this  handle  engages  with  the  steps  D,  which  gradually  rise  to  a  higher 
plane,  so  that  by  turning  the  handle  in  one  direction  the  tank  valve  will  be  raised, 
and  turning  it  in  the  opposite  direction  the  tank  valve  will  be  lowered  until  finally  it 
shuts  off  the  feed  water  from  the  engine. 


IKON  TENDER-FRAME. 

564.  Figs.  962  to  964  show  a  good  construction  of  an  iron  tender-frame.  It 
consists  of  four  channel  irons  A  A,  10  inches  deep  by  2f  inches  wide.  These  channel 
irons  are  placed  longitudinally,  and  their  ends  are  riveted  by  wrought-iron  knee  plates 
to  the  transverse  end  plates  B  of  wrought-iron,  which  are  1  inch  thick.  The  ends  of 
the  channel  irons  are  also  held  together  by  the  plates  C  C,  £  inch  thick,  riveted  to  the 
bottom  flange  of  the  channel  irons,  and  further  stiffness  is  secured  by  the  braces  G  G. 
To  the  plates  C  C  the  cast-iron  draw-heads  are  bolted,  which,  as  will  be  seen,  differ  in 
design  from  those  shown  on  the  frame  in  Fig.  943.  The  hole  D  near  the  front  end 
takes  the  coupling  pin,  and  the  holes  E  E  on  each  side  of  D  take  the  pins  for  Un- 
safely chain.  Oak  bumpers  are  bolted  to  the  transverse  plates  B  B  as  shown.  The 


J.^ 


MOIH-:I;\ 


571 


cast -iron  sockets  P  P  for  the  truck  center-pin  are  bolted  to  the  transverse  bars  F  JP, 
which  are  riveted  to  the  channel  irons. 

BRAKE   GEAR. 

565.  In  the  plan  of  this  frame  a  very  efficient  brake  gear  is  shown.  The  stand  8 
supports  the  upright  brake  shaft  (not  shown  here)  around  which  the  chain  S2  is 
wound;  this  chain  is  connected  to  the  rod  AT,  which  pulls  on  the  lever  I;  this  lever 
is  fulcrumed  at  its  center.  A  small  chain  wheel  works  on  the  fulcrum  pin,  and  two 
wheels  M  M  of  similar  form  work  on  the  pivots  fastened  to  the  lever  at  a  short 
distance  from  its  ends.  A  chain  L  L  passes  around  these  wheels  as  shown;  one  end 
of  this  chain  is  connected  to  the  rod  Nt  which  works  the  brakes  on  the  front  truck,  and 
the  other  end  of  the  chain  is  fastened  to  a  rod  on  the  opposite  end  of  lever  which  works 
the  brakes  on  tho  rear  truck.  With  this  arrangement  the  brakes  are  put  in  action 
by  hand,  and  the  brake  pressure  on  the  front  and  rear  truck  is  equalized  by  the  chain 
L  L.  If  it  is  desirable  to  put  the  brakes  in  action  by  air  pressure,  the  end  of  the 
lever  J  which  engages  with  the  rod  N  is  connected  to  the  piston  of  the  air  cylinder 
without  disturbing  any  of  the  other  parts  of  the  brake  gear  already  described.  With 
this  arrangement  the  brakes  can  be  put  into  action  either  by  hand  or  by  air  pressure. 


TENDER  TRUCKS. 

566.  There  are  quite  a  variety  of  tender  tracks,  but  they  all  have  one  feature  in 
common  which  differs  from  the  engine  truck,  namely,  all  the  axle  journals  are  placed  on 

the  outside  of  the  wheels, 
which  is  not  admissible 
on  engine  trucks,  because 
their  frames  must  be  kept 
clear  of  the  cylinders.  In 
placing  the  journals  out- 
side of  the  wheels  their 
bearings  are  more  acces- 


sible, and  the  boxes  can 

be  entirely  closed  over  the  ends  of  axle,  thereby 
preventing,  to  a  great  extent,  dust  from  entering 
the  boxes,  and  preventing  oil  from  leaking  out  at 
the  ends.  Also  in  this  position  the  journal  bearings 
can  be  removed  and  replaced  by  new  ones,  and 
the  boxes  can  be  repacked  with  waste  with  much 
less  difficulty  than  when  the  journals  are  on  the 
inside  of  the  wheels. 

Figs.  965  to  967  show  one  of  the  simplest  constructions  of  tender  tracks.  It 
consists  of  a  wooden  bolster  B  (generally  made  of  oak),  which  is  supported  by  the 
wrought-iron  bars  C,  C2,  (7:!,  and  C4.  These  bars  are  bolted  direct  to  the  axle-boxes 
A  A.  In  this  class  of  tracks  the  springs  are  simply  laid  on  top  of  the  truck  frame. 


Fig!  966. 


572 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


For  the  rear  truck  here  shown,  which  is  a  side-bearing  truck,  four  springs  are  used ; 
two  of  these  springs  (one  on  each  side  of  truck)  extend  from  the  center  of  one  axle  to 
the  center  of  the  other  one.  The  ends  of  these  springs  rest  in  the  cast-iron  pockets 
I)  Z>,  which  are  bolted  over  each  axle-box  to  the  frame.  Another  pair  of  springs  extend 
across  the  truck,  one  on  each  side  of  the  center  pin  socket  F;  their  ends  rest  in  the 

the  pockets  e  e,  which  are 
cast  in  the  bearings  E  E. 
The  centers  of  these  four 
springs  support  the  rear 
end  of  the  tender  frame. 

_^'  MB  '  ftM         >W      VW     — 5?»HN =i 


The  front  track,  which 
is  similar  in  construction, 
is  shown  in  Figs.  968  to 
970.     This  truck  is  a  center-bearing  truck  and 
consequently  it  has  no  outside  springs.     It  has 
only  two  springs  which  extend  across  the  truck, 
one  on  each  side  of  the  center-pin  socket  F; 
the  ends  of  these  springs  rest  in  the  pockets 
e  e  cast  in  the  bearings  E  E;   the  center  of 
these  springs  support  the  front  end  of  tender 
Fig.  970.  frame. 

Figs.  971  to  974  show  different  views   and 

sections  of  another  type  of  tender  truck.  In  this  one  the  framing  consists  of  the 
wrought-iron  bars  C,  C2,  and  (73,  which  are  bolted  to  the  axle-boxes  A  A.  The 
lower  bolster  B2  is  supported  by  the  bars  C2  and  (73;  on  this  bolster  four  springs 
S  S  are  laid,  whose  centers  support  the  upper  bolster  B,  which  is  free  to  move  up 
and  down  between  the  guides  H  H.  For  the  rear  truck  two  side  bearings  E  are 
bolted  to  the  top  face  of  the  upper  bolster  B,  on  which  the  rear  end  of  the  tender  frame 
rests.  These  bearings  make  this  truck  a  side-bearing  one.  For  the  front  truck  the 
bearings  E  are  not  used ;  the  front  end  of  the  tender  frame  .is  supported  by  the  center- 
pin  socket  F. 

Figs.  975  to  977  show  different  views  of  a  truck  used  for  the  tender  shown 
in  Fig.  935.  The  frame  F  is  made  of  wrought-iron,  the  pedestals  A  of  cast-iron.  In 
this  case  the  brake  beams  B  are  often  made  of  ash  or  oak ;  for  other  tenders  they 
are  sometimes  of  wrought-iron  built  up  in  the  form  of  a  truss.  The  axle  guards  C  are 
simply  wrought-iron  straps  and  bent  in  the  form  of  a  U.  Otherwise  the  construction 
of  the  truck  is  so  plainly  shown  that  a  further  description  is  unnecessary. 


TENDERS   WITH   WATEE-SCOOPS. 

567.  The  tanks  for  tenders  heretofore  shown  are  filled  by  means  of  a  hose  or  pipe 
attached  to  the  water  tanks  or  hydrants  in  water  stations.  This  often  necessitates 
stops  at  comparatively  short  intervals,  and  consequently  causes  delays  undesirable  for 
express  trains.  In  order  to  avoid  such  delays  as  much  as  possible  some  tenders  are 
provided  with  water-scoops  for  taking  water  while  the  engine  is  running,  the  water 


573 


574 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


being  taken  from  a  long  narrow  trough  laid  between  tlie  rails.     The  whole  arrange- 
ment is  shown  in  Figs.  978,  979,  980,  and  980A. 

The  tank  differs  from  the  ordinary  one  by  having  a  pipe  A,  Fig.  978,  running 
through  it  leading  the  water  up  to  the  inside  of  the  manhole,  which  is  shown  in  dotted 
lines,  from  whence  it  is  discharged  into  the  tank.  The  smaller  manhole,  shown  in  full 
lines,  simply  indicates  one  as  is  used  in  the  ordinary  tank  without  the  pipe  A.  The 


Fig.  976 


AM. BANK  NOTE-CO. 


pipe  B  in  Fig.  979  is  a  continuation  of  the  pipe  A,  a  water-tight  joint  being  made 
between  them.  The  scoop  C  is  hinged  to  the  pipe  B  so  that  the  scoop  can  be  raised  or 
lowered  by  means  of  the  system  of  levers  shown  plainly  in  Figs.  979,  980,  and  980A ; 
the  raising  and  lowering  being  accomplished  from  the  floor  of  the  tender.  The  scoop 
is  shown  in  its  lowest  position  with  its  lower  end  dipping  a  few  inches  into  the  water 
in  the  trough.  The  bottom  of  the  scoop  acts  like  uu  inclined  plane  and  takes  advau- 


.MOI>/:/;\ 


575 


tage  of  the  inertia  of  the  water,  which,  being  at  rest  in  the  trough  and  the  engine  run- 
ning at  a  rate  of  30  to  40  miles  an  hour,  enables  the  scoop  to  lift  the  water  through 
the  pipe  A,  and  is  finally  discharged  into  the  tank.  The  height  through  which  the 
water  can  be  lifted  will  greatly  depend  on  the 
speed  ot  the  train. 

TENDER-TRUCK  AXLE-BOXES 

568.  Fig.  981  shows  a  longitudinal  section  of 
an  ordinary  tender- truck  axle-box.  Fig.  982  shows 
a  horizontal  section,  Figs.  983  and  984  show  re- 
spectively an  outside  end  view  and  plan  of  the 
same  box.     It  consists  of  a  cast-iron  casing  A,  a 
brass  bearing  B,  and  a  cast-iron  wedge  C.     This 
axle-box  is  suitable  for  a  truck  as  shown  in  Fig. 
972,  and  it  is  rigidly  bolted  to  the  frame  by  two 
turned  bolts  driven  through  the  holes  Z>,  which 
are  bored  out  at  the  ends  only,  and  cored  in  the 
center  somewhat  larger  than  the  diameter  of  the 
bolts.     The  brass  B  rests  on  the  journal  J,  and  is 
held  in  position  sideways  by  the  projections  E  E 
(Fig.  983)  cast  to  the  inner  side  of  the  casing  A. 

569.  The  wedge  C  is  placed  in  the  box  after 
the  brass  has  been  placed  in  position,  and  its  pur- 
pose is  simply  to  fill  the  space  required  for  slipping 
the  brass  over  the  collar  of  the  journal. 

570.  Many  brasses  have  a  convex  back  so  as  to 
give  them  only  a  small  area  of  con- 
tact with  the  wedge,  thereby  giv- 
ing freedom  to  the  axle  to  adjust 

it -i 4f  to  any  unevenness  of  the 
track.  On  account  of  this  small 
area  of  contact  the  brass  must  be 
made  comparatively  thick  so  that 
its  ends  will  not  spring  upwards 
and  wear  the  journal  more  at  its 
center  than  at  its  ends. 

The  brasses  are  bored,  having 
frequently  a  radius  from  .}.,  to  ^ 
inch  greater  than  that  of  the  jour- 
nal. The  edges  /•'  F  of  the  brass 
(Fig.  983)  should  be  well  rounded 
so  as  to  prevent  them  from  scrap- 
ing the  oil  off  the  journal.  The 
width  of  the  bearing  surface  should 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


MOI>Kl;\    l.ornHuril  K    COXSTIit'CTlOX. 


577 


extend  about  one-third  around  the  circumference  of  the  journal,  and  the  length  of  the 
brass  is  usually  i  inch  less  than  that  of  the  journal. 

571.  Sometimes  the  bearing  surface  of  the  brass  is  composed  of  a  lead  lining, 
which,  it  is  believed,  prevents  heat- 
ing to  a  considerable  extent,  par- 
ticularly when  new  brasses  are 
used  on  worn  journals.  That  lead 
lining  has  given  good  results  can- 
not be  denied,  but  reliable  and 
definite  values  are  lacking.  This 
lining  should  not  be  more  than 
i*e  inch  thick;  if  it  is  more  than 
this,  the  lead  will  be  forced  out  at 
the  edges  so  as  to  prevent  a  free 


flow  of  oil  between  the  wearing  sur- 
faces, and  otherwise  give  trouble. 


Fig.  980  A. 


572.  To  prepare  the  brass  for  the  lining  it  is  first  bored,  then  well  heated  over 
a  coke  fire,  then  cleaned  with  muriatic  acid  and  tinned.    After  this  it  is  placed  at  a 


578 


MODERN  LOCOMOTIVE  CO\STKUCTJOJf. 


Fig.  985. 


correct  distance  from  a  mandrel  of  suitable  diameter,  and  warmed  again  to  enable  the 

lead,  when  poured  between  the  mandrel  and  brass,  to  adhere  to  the  tin.     This  process 

was  patented  by  Mr.  Hopkins,  but  the  patent  has  now  expired. 

573.  In  designing  the  foregoing  class  of  axle-box,  care  should  be  taken  to  have  the 

centers  of  the  holes  D  in  line  with  the 

center  of  the  journal  as  shown. 

The  bottom  of  the  axle-box  is  filled 

with  cotton  or  woolen  waste  saturated 

with  oil,  which  bears  against  the  bottom 

.  of  the  journal  and  keeps  it  lubricated. 
The  rib  G  cast  to  the  bottom  of  the 

box,  Fig.  981,  prevents  the  waste  from 

being  forced  too  far  back.    This  rib  is  not  used  in  all  boxes ;  and  when  not  used,  the 

waste  touches  the  end  of  the  box.     The  opening  H  takes  the  dust  guard  either  made 

of  wood  or  leather.     The  dust  guard  is  made  to  fit  the  axle  at  I,  and  prevents  the 

dust  from  getting  in  and  the 
oil  from  leaking  out. 

574.  The  front  cover, 
which  is  not  shown,  is  in 
this  design  of  box  fastened 
to  it  by  two  screw  bolts. 
Trouble  is  sometimes  occa- 
sioned by  the  loss  of  the 
screw  bolts  caused  by  care- 
lessness and  neglect  to  screw 
them  in  tight ;  consequently, 
on  some  roads  the  cover  or 
lid  is  made  in  the  form  of  a 
wedge,  as  shown  in  Fig.  985. 
On  other  roads  the  cover  has 
a  lug  cast  to  each  side  which 
fit  in  tapered  grooves  A  A 
cast  to  the  sides  of  the  box, 
as  shown  in  Fig.  986.  This 
cover  wedges  itself  tightly  in 
position  when  dropped  into 
its  place.  The  grooves  are 
so  arranged  as  to  permit  the 
box  to  be  readily  opened,  but 

the  cover  cannot  be  conveniently  removed  wholly,  and  therefore  the  liability  of  losing 

it  is  much  reduced,  while  the  annoyance  of  dust  flying  into  the  box  is  to  a  great  extent 

avoided. 

575.  The  axle-box  shown  in  Fig.  986  is  similar  in  design  to  that  shown  in  Fig. 

981,  but  it  has  no  wedge,  and  the  dust  guard  enters  at  the  top  of  the  box  instead  of 


Eig.986. 


579 


entering  at  the  bottom ;  the  bolts  which  secure  the  box  to  the  frame  lie  in  grooves 
cast  in  the  sides  of  the  box  instead  of  being  driven  into  the  holes.  A  very  small 
amount  of  clearance  is  allowed  for  the  bolts  in  the  grooves.  This  is  done  not  only  for 
cheapness,  but  also  to  provide  easy  means  for  taking  the  bolts  out  when  it  becomes 
necessary  to  remove  the  box,  thereby  facilitating  repairs. 

576.  Fig.  987  shows  another  box  similar  in  design  to  that  shown  in  Fig.  981,  but 
with  this  difference :  the  sides  of  the  box  are  arranged  for  a  pedestal  in  which  the  box 
is  free  to  move  up  or  down.    A  box  of 

this  kind  is  suitable  for  a  truck  shown 
in  Fig.  976.  The  pedestal  must,  of 
course,  be  placed  central  with  the  axle 
journal.  It  will  also  be  noticed  that  the 
cover  of  this  box  is  secured  by  one 
screw  bolt  only. 

577.  The  design  of  box  shown  in 
Fig.  988  explains  itself.     It  is  made 

for  an  axle  journal  which  has  no  end  collar.  The  manner  of  holding  its  cover 
in  position  is  somewhat  peculiar;  it  is  held  by  a  spiral  spring,  and  in  order  to 
open  the  box  the  cover  must  be  sprung  beyond  the  projections  cast  to  the  end  of 
the  box. 

578.  There  are  many  other  types  of  boxes  used,  their  design  depending  chiefly  on 
the  form  of  brasses  adopted,  the  manner  of  fastening  the  covers,  and  the  type  of  truck 


Fig.  988. 


for  which  they  are  intended.    But  the  features  which  enter  into  the  design  of  the 
tender-truck  boxes  are  shown  in  the  foregoing  illustrations. 

Fig.  988A  shows  two  views  of  a  tender-truck  axle-box,  and  Fig.  988B  shows 
different  views  of  its  bearing  and  wedge.  This  box  was  adopted  as  a  standard  by  the 
American  Railway  Master-Mechanics'  Association. 


580 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


TENDER-TKTJCK  AXLES. 


579.  In  designing  a  tender-truck  axle  the  first  problem  which  presents  itself  is  to 
find  the  dimensions  of  the  journals.  These  must  not  only  be  strong  enough  to  hold 
the  weight  of  the  tender,  fuel,  and  water,  but  they  must  also  be  so  proportioned  as  not 


CTION  OK  CENTRE  LINE  OF  AXl 


rH  "_ 


TRANSVERSE  SECTION  OF  KEY  oTTtlNE  A  B       j 


--« — I >) 

._*»''. 


LONGITUDINAL  SECTION  OF  KEY  AND  BEARING 

[,L|i_  OUTSIDE  END  VIE*  OF  BEARING  AND  KEY 


TRANSVERSE  SECTION  OF  BEARING  ON  LINE  0  0 

'A..n«k»«.fcii.T.  Fig,  988.  B. 


to  heat;  and  generally,  when  they  meet  the  requirements  of  the  latter  condition,  they 
will  also  be  strong  enough  to  hold  the  weight  placed  upon  them.  To  prevent  heating, 
the  pressure  upon  the  journal  must  not  exceed  a  certain  limit. 


MODERN  LOCOMOTITE  CONSTRUCTION. 


581 


Referring  to  Fig.  983,  we  notice  that  the  brass  touches  only  about  one- third  of  the 
circumference  of  the  journal ;  but  for  the  sake  of  simplicity  in  making  the  computa- 
tions it  is  usually  assumed  that  the  brass  covers  one-half  of  the  circumference,  so  that 
pressure  may  be  referred  to  the  projected  area  of  the  journal,  which,  as  we  have  seen, 
is  equal  to  the  product  of  the  diameter  into  the  length  of  the  journal. 

580.  To  obtain  the  weight  on  each  journal  we  divide  the  total  weight  of  tender, 
fuel,  and  water,  minus  the  weight  of  wheels  and  axles,  by  the  number  of  journals  under 
the  tender ;  the  quotient  will  be  the  weight  which  each  journal  has  to  support.     Divid- 
ing this  weight  by  the  projected  area,  we  obtain  the  pressure  per  square  inch  of 
projected  area.    This  pressure  varies  under  different  tenders  from  170  to  350  pounds. 
We  believe  that  the  safest  practice  is  not  to  allow  it  to  exceed  225  pounds  per  square 
inch  for  a  joiirnal-surface  speed  of  5  feet  per  second. 

581.  On  many  roads  the  tender-truck  journals  are  3J  inches  diameter  and  7  inches 
long  for  tenders  weighing,  when  empty,  from  20,000  to  24,000  pounds,  and  with  fuel 


1        ""       1    nt"- 

.  -  TlA  / 

lili 

-i'  rr                   -* 

you. 

i 

D 

W 

*, 

i                     V 

7 

-1 

E 

-• 

? 

m 

T 

Li-rnrtb-ovtT-alK!-ftr-lli.-iMi 

FINISHED     SIZES. 

Master  Car  Builders'  Standard  Axle  for  30,000  pound  Cars.     Master  Mechanics'  .Standard  Axle  for  Light  Tenders. 


*v 

g 

>* 

is, 

~1T~                          ~7      -T      T-  ,t     .I"' 

^P                                                                                                           V                     J?                   ^               *9      1            •* 

f                        •  ?     *     •?    f  !  * 

.-*• 

5«: 
•* 

,4 

-V 

FINISHED     SIZES. 

Master  Car  Builders'  Standard  Axle  for  60,000  pound  Cars.     Master  Mechanics'  Standard  Axle  for  Heavy  Tenders. 

and  water  weighing  from  41,000  to  58,000  pounds.  For  tenders  weighing,  when 
empty,  from  27,000  to  30,000  pounds,  and  with  fuel  and  water  weighing  from  60,000 
to  74,000  pounds,  the  journals  are  generally  4  inches  diameter  and  8  inches  long. 
Other  roads  have  adopted  the  Master  Car  Builders'  standard  for  the  tenders  so  as  to 
avoid  carrying  extra  sizes  of  axles  for  tenders  in  stock.  These  standards  are  illus- 
trated in  Figs.  989  and  990;  they  are  recommended  by  the  Master-Mechanics 
Association. 

The  axle  shown  in  Fig.  989  has  journals  3$  inches  in  diameter  and  7  inches  long. 
It  is  suitable  for  tenders  whose  total  weight  varies  from  40,000  to  58,000  pounds. 
The  other  axle  shown  in  Fig.  990  has  journals  4£  inches  diameter,  8  inches  long. 
It  is  suitable  for  tenders  weighing  in  complete  working  order  from  60,000  pounds  and 
upwards.  These  axles  are  made  either  of  iron  or  steel. 


582  MODERN   LOCOMOTIVE    CONSTRUCTION. 

582.  It  will  be  noticed  that  the  pressure  per  square  inch  of  projected  area  of  the 
tender-truck  axle  is  greater  than  that  of  the  engine-truck  axles.     This  is  admissible, 
because  the  tender-truck  wheels  are  usually  larger  in  diameter  than  the  engine-truck 
wheels.    Again,  in  engine-trucks  the  journals  are  of  necessity  placed  inside  of  the 
wheels,  and  therefore  they  have  to  be  comparatively  large  in  diameter  and  limited  lengths 
so  as  to  give  the  required  strength  to  the  axles ;  these  conditions  give  a  greater  surface 
velocity  to  the  engine-truck  journals  than  that  of  the  tender-truck  journals,  and 
consequently  the  pressure  per  square  inch  of  projected  area  of  the  former  journals 
cannot  be  as  large  as  the  pressure  on  the  latter.     The  journals  in  tender-trucks  are 
generally  better  protected  from  dust,  and  the  load  on  them  decreases  as  the  fuel  and 
water  is  used  up,  all  of  which  tends  to  lessen  the  danger  of  heating. 

583.  To  find  the  dimensions  of  journals  for  tenders  weighing  less  than  given  in 
Art.  581,  we  subtract  the  weight  of  the  wheels  and  axles  from  the  total  weight  of  the 
tender,  water,  and  fuel,  and  divide  the  remainder  by  225 ;  the  quotient  will  give  the 
sum  of  the  projected  areas  of  the  journals.    Dividing  the  latter  quotient  by  the  number 
of  journals,  we  obtain  the  projected  area  of  each ;  if  now  the  length  is  known,  we  can 
easily  find  the  diameter.     But  for  the  sake  of  simplicity  we  may  at  once  divide  the 
total  weight  of  tender,  water,  and  fuel  by  250,  and  divide  this  quotient  by  the  number 
of  journals,  which  will  give  us  the  projected  area  of  each  near  enough  for  practical 
purposes,  from  which  the  diameter  and  length  of  journal  can  be  easily  found  if  we 
know  the  ratio  between  the  two. 

EXAMPLE  151. — The  total  weight  of  the  tender,  water,  and  fuel  is  38,000  pounds ; 
it  has  two  four-wheeled  trucks ;  the  length  of  the  journal  is  to  be  equal  to  twice  its 
diameter ;  find  the  length  and  diameter  of  the  journals. 


Here  we  have  38000 

250 


=  152  square  inches 


for  the  projected  area  of  eight  journals,  and 

152 

-5-.  =  19  square  inches 

o 

for  the  projected  area  of  each  journal. 

The  ratio  between  the  length  and  diameter  of  journal  is  2 ;  hence  the  diameter 

/19T 
can  be  found  by  the  rule  given  in  connection  with  engine-trucks,  thus :  V  -5-  =  3.08, 

V        _ 

say  3  inches  for  the  diameter  of  the  journal  and  3x2  =  6  inches  for  its  length. 

584.  Although  journals  are  generally  strong  enough  to  support  the  weight  placed 
upon  them  when  they  are  correctly  designed  to  prevent  heating,  the  question  may 
arise :  What  load  can  a  journal  support  ? 

The  worst  condition  to  which  a  journal  may  be  subjected  is  when  the  brass  bears 
only  on  its  outer  end.  In  cases  of  this  kind  we  may  consider  the  journal  to  be  a  beam 
supported  at  one  end  and  loaded  at  the  other.  For  computing  the  load  which  a  beam 
of  this  form  can  support  with  safety,  we  have  the  following  formula : 

S  x  3.1416  x  d3 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


583 


in  which  L  denotes  the  load  in  pounds ;  S,  the  stress  per  square  inch  on  the  fibers 
most  remote  from  the  neutral  surface;  for  wrought-iron  the  value  of  S  is  generally 
taken  at  12,000  pounds,  and  for  steel,  16,000  pounds ;  d  denotes  the  diameter  in  inches ; 
and  /,  the  length  of  the  journal  in  inches.  To  show  the  application  of  this  formula,  we 
will  take  the  following  example : 

EXAMPLE  152. — What  load  can  a  wrought-iron  journal  safely  support,  its  diameter 
being  3J  inches  and  its  length  7  inches,  under  the  assumption  that  the  whole  load  is 
concentrated  at  the  end  of  the  journal ! 

Substituting  in  the  formula  for  the  symbols  their  values,  we  have, 


L  = 


12000  x  3.1416  x  3.75 3 


=  8875  pounds. 


32  x  7 

If  this  size  of  journal  is  used  under  a  tender  weighing  58,000  pounds,  then  the  actual 
weight  on  each  journal  will  not  reach  7,250 
pounds,  which  shows  us  that  the  journal  is 
strong  enough  for  the  work  it  has  to  do. 

If  we  consider  the  journal  to  be  a  beam 
supported  at  one  end  and  uniformly  loaded, 
then  the  load  which  it  can  support  is  com- 
puted by  the  following  formula : 

8  x  3.1416  x  d3 

in  which  the  symbols  denote  the  same  quan- 
tities as  before.  Working  out  Ex- 
ample 152  by  this  formula,  we  find 
that  this  journal  can  support  a  load 
uniformly  distributed  of  17,750 
pounds,  which  shows  us  that  the 
journal  is  more  than  strong  enough 
for  a  tender  whose  total  weight  is 
58,000  pounds. 

Formula  (a)  can  be  used  for 
finding  the  load  which  can  be  sup- 
ported by  any  beam  of  circular 
cross-section  fixed  at  one  end  and 
loaded  at  the  other. 

Formula  (b)  can  be  used  for 
finding  the  uniformly  distributed 
load  which  can  bo  supported  by 
any  beam  of  circular  cross-section 
and  fixed  at  one  end. 

TRUCK  WHEELS. 


585.  Of  the  iiKiiiy  types  of  wheels  now  in  use  under  cars  and  tenders,  the  cast-iron 
.wheel  with  chilled  treads  is  the  most   common  one.     A  section  of  this  wheel  is  shown 


584 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


in  Fig.  991 ;  its  form  should  be  such  that  no  part  of  the  wheel  is  unduly  strained 
during  cooling  after  it  has  been  cast.  The  section  in  Fig.  991  is  of  a  type  used  under 
cars  and  tenders  on  the  Pennsylvania  R.  R. 

586.  For  tender-trucks  the  face  a  of  the  hub  does  not  reach  the  outer  face  c  of  the 


Fig.  992. 


tread  so  as  to  allow  sufficient  space  for  the  axle-box ;  and  in  order  to  obtain  the  neces- 
sary depth  for  the  hub  its  inside  face  b  projects  beyond  the  face  of  the  flange  d.  In 
this  respect  the  tender-truck  wheel  differs  from  the  engine-truck  wheel ;  in  the  latter 


Pig.  993. 


the  hub  is  placed  longitudinally,  nearly  central  with  the  tread,  so  as  to  allow  space  for 
the  axle-box,  which  in  the  engine-truck  is  placed  on  the  inside  of  the  wheel. 

587.  Fig.  992  shows  a  type  of  wheel  manufactured  by  the  Allan  Paper  Wheel  Com- 
pany, of  Chicago,  111.     In  this  type  the  tire  is  shrunk  on  a  cast-iron  center  and  held  by 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


585 


two  retaining  rings.    The  form  of  these  rings  and  that  of  the  tire  are  such  that,  should 
the  latter  break  across,  the  retaining  rings  will  hold  the  broken  tire  in  place. 

Fig.  993  shows  a  paper  wheel  also  manufactured  by  the  Allan  Paper  Wheel  Com- 
pany.   In  this  wheel  the  tire  is  joined  to 
the  hub  by  two  wrought-iron  plates. 

588.  Fig.  994  shows  a  steel  wheel  manu- 
factured by  the  Boies  Steel  Wheel  Company, 
of  Scranton,  Pa.     Its  construction  is  very 
simple.     The  wheel  is  composed  of  four 
principal   parts:    the   steel    tire,  the  two 
corrugated    steel    plates,   and  a    cast-iron 
hub.    The  tire  is  shrunk  on  the  corrugated 
plates,  and  the  whole  is  firmly  bolted  to- 
gether. 

589.  An  excellent  type  of  truck  wheel 
is  made  at  the  Baldwin  Locomotive  Works, 
Philadelphia.      The    tire   is    placed    on    a 
wrought-iron  center  of   the  spoke   type. 
The  centers    are  made   by  an   improved 
method  patented  by  Samuel  M.  Vauclain, 
Superintendent  of  the  Baldwin  Locomotive 
Works.     These  wheels  are  veiy  strong  and 

have  a  fine  finish,  resembling  a  well-made  cast-iron  wheel  having  spokes  of  elliptical 
cross-section. 

GENERAL  TYPES  OF  LOCOMOTIVES. 

590.  Fig.  995  shows  a  longitudinal  section  and  a  half  plan  of  the  machinery  of  an 
eight-wheeled  passenger  locomotive  built  by  the  Grant  Locomotive  Works,  formerly  at 
Paterson,  N.  J.,  now  at  Chicago,  111.     The  right-hand  side  of  Fig.  996  shows  the  rear 
end,  and  the  left-hand  side  the  front  end,  of  the  same  engine  with  the  smoke-box  front 
and  bumper  removed.     The  left-hand  side  of  Fig.  997  shows  a  transverse  section  taken 
in  front  of  the  guide-yoke,  and  the  right-hand  side  shows  a  transverse  section  taken 
behind  the  sand-box.   This  engine  has  a  short  smoke-box,  and  is  equipped  with  pumps, 
whose  location,  with  the  arrangement  of  suction  and  delivery  pipes  such  as  are  gen- 
erally adopted  in  this  class  of  engine,  is  plainly  shown. 

591.  Fig.  998  shows  an  elevation  of  another  eight-wheeled  passenger  engine  and 
a  half  plan  of  its  machinery.    This  is  the  staudai'd  design  adopted  by  the  Pennsylvania 
R.  R.    In  these  illustrations  is  shown  the  manner  of  attaching  the  Westinghouse  brake. 
The  engine-truck  is  shown  in  Figs.  886A,  886B,  and  886C.     This  engine  is  doing  excel- 
lent service.     Its  principal  dimensions  are  as  follows : 


SPECIFICATIONS    FOB    STANDARD    P.  R.   R.   LOCOMOTIVE   WITH    TENDER. 

DRIVERS — 160   POUNDS   STEAM   PRESSURE. 


CLASS    "P" — 08-INCH 


Gauge 4  ft.  9  in. 

Number  of  Pairs  of  Driving  Wheels 2. 

Diameter  of  Driving  Wheels  68  in. 

Wheel  Centers Cast-iron. 


Tires Steel. 

Total  Wheel  Base 22  ft.  8$  in. 

Length  of  Rigid  Wliccl  Base 7  ft.  '.!  in. 

Diameter  of  Driving-axle  Bearing Sin. 


_L  _  zr=D 


T 


580 


bl 


587 


c  H 

Sm 


588 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


589 


Length  of  Driving-axle  Bearing 10|  in. 

Diameter  of  Main  Crank-pin  Bearing 4J  in. 

Length  of  Main  Crank-pin  bearing 4|  in. 


Diameter  of  Parallel-rod  Bearing  > 


3}  in. 


Length  of  Parallel-rod  Bearing     f  Front  and  Back  3J  in. 

Number  of  Wheels  in  Front  Tnick 4. 

Diameter  of  Wheels  in  Front  Truck 33  in. 

Material  of  Wheels  in  Front  Truck,  Cast-iron  Steel  Tired. 

Diameter  of  Truck-axle  Bearing 5$  in. 

Length  of  Truck-axle  Bearing 10  in. 

Type  of  Truck Rigid  Center. 

Cylinders  and  Steam-chests Outside. 

Spread  of  Cylinders 77  in. 

Diameter  of  Cylinders 184  in. 

Length  of  Stroke  24  in. 

Position  of  Valve  Gear Between  Frames. 

Type  of  Valve  Gear Shifting-link  Motion. 

Travel  of  Valve 5  in. 

Outside  Lap  of  Valve f  in. 

Inside  Lap  of  Valve None. 

Lead  of  Valve -fg  in. 

Throw  of  Eccentric 5  in. 

Length  of  Steam  Ports 17J  in. 

Width  of  Steam  Ports li  in. 

Width  of  Exhaust  Port 2i  in. 

Kind  of  Frames Wrought-iron,  Inside. 

Distance  between  Centers  of  Frames 44  in. 

Boiler  Material Steel. 

Thickness  of  Boiler  Sheets,  Dome f  in. 

Thickness  of  Boiler  Sheets,  Barrel  and  Smoke-box.  -fe  in. 

Thickness  of  Boiler  Sheets,  Outside  Fire-box fg  in. 

Maximum  Internal  Diameter  of  Boiler  >  .       56J  in. 

Minimum  Internal  Diameter  of  Boiler  }  53$  in. 

Height  to  Center  of  Boiler,  from  Top  of  Bail 864  in. 

Number  of  Tubes 210. 

Inside  Diameter  of  Tubes 1J  in. 

Outside  Diameter  of  Tubes 2  in. 

Tube  Material Wrought-iron. 

Length  of  Tubes  between  Tube  Sheets 136  in. 

External  Heating  Surface  of  Tubes 1244  sq.  ft. 

Fire  Area  through  Tubes 3.5  sq.  ft. 


Length  of  Fire-box  (Inside) 9  ft.  11J  in. 

Width  of  Fire-box  (Inside) 3  ft.  4  in. 

Height  of  Crown  Sheet  above  Top  of  Grate  (Center 

of  Fire-box) 3  ft.  Si  in. 

Inside  Fire-box  Material Steel. 

Thickness  of  Inside  Fire-box  Sheets,  Sides A  >n- 

Thickness  of  Inside  Fire-box,  Crown |  in. 

Thickness  of  Inside  Fire-box,  Back f  in. 

Thickness  of  Tube  Sheets,  Front 4  in. 

Thickness  of  Tube  Sheets,  Back f  in. 

Tube  Sheet  Material Steel. 

Heating  Surface  of  Fire-box 138  sq.  ft. 

Total  Heating  Surface 1382  sq.  ft. 

Fire  Grate  Area 33.25  sq.  ft. 

Diameter  of  Smoke-stack  (Straight) 18  in. 

Height  of  Stack  above  Top  of  Kail 15  ft.  0  in. 

Width  of  Cab  Roof 9  ft.  6  in. 

Height  of  Cab  Roof  from  Rail 12  ft.  10*  in. 

Width  of  Cab 9  ft. 

Size  of  Exhaust  Nozzle 2f  x  3i  in. 

Pressure  of  Steam  per  Square  Inch 160  Ibs. 

Nature  of  Fuel Anthracite  Coal. 

Weight  of  Engine  Empty 96, 100  Ibs. 

Weight  on  Drivers 65,150  Ibs. 

Weight  on  Truck 30,950  Ibs. 

Weight  of  Engine  in  Working  Order 106,500  Ibs. 

Weight  on  First  Pair  of  Drivers 36,500  Ibs. 

Weight  on  Second  Pair  of  Drivers 36,850  Ibs. 

Weight  on  Truck 33,150  Ibs. 

Engine  fitted  with  Driver  Brake. 

Capacity  of  Tank 3000  gal. 

Capacity  of  Coal-box 9900  Ibs. 

Number  of  Wheels  under  Tender 8. 

Diameter  of  Wheels  under  Tender 33  in. 

Material  of,"Wheels  under  Tender,  Cast-iron  Chilled  Tread. 

Diameter  of  Tender-truck  Journals 4  in. 

Length  of  Tender-truck 8  in. 

Weight  of  Tender,  Empty 29,800  Ibs. 

Weight  of  Tender,  Loaded 65,500  Ibs. 

Tender  fitted  with  Water-scoop. 


It  will  be  noticed  that  this  engine  has  an  extension  smoke-box,  and  the  driving- 
wheel  springs  are  placed  below  the  axle-boxes  instead  of  above  them,  as  shown  in 
Fig.  995. 

592.  Fig.  999  shows  an  elevation,  and  Figs.  1000,  1001,  and  1002  show  end  views 
and  section  of  a  ten-wheeled  engine  built  by  the  Baldwin  Locomotive  Works,  Phila- 
delphia, Pa. 

Fig.  1003  shows  an  elevation  of  a  consolidation  engine  also  built  by  the  Baldwin 
Locomotive  Works. 

593.  Fig.   1004  shows  a   powerful  six-wheeled  switching  engine  built  by  the 
Pennsylvania  R.  E.     Its  principal  dimensions  are  as  follows : 

SPECIFICATIONS   FOR  STANDARD  P.   R.   R.   CLASS   "  M "   SHIFTING  ENGINE  WITH  TENDER. 


Gauge 4  ft.  9  in. 

Number  of  Pairs  of  Driving  Wheels 3. 

Diameter  of  Driving  Wheels 50  in. 

Wheel  Centers Cast-iron. 


Tires Steel. 

Total  Wheel  Base 10  ft.  8  in. 

Length  of  Rigid  Wheel  Base 10  ft.  8  in. 

Diameter  of  Driving-axle  Bearing 7J  in. 


590 


MODEKN  LOCOMOTIVE   CONSTRUCTION. 


3 


05 


MODEHX  LOCOMOTirK  CONSTRUCTION. 


591 


•  8 

!— I 

<3> 

.  £ 


592 


593 


594 


MODERN  LOCOMOTIVE   COXSTBUCTION. 


Length  of  Driving-axle  Bearing  .................  7|  in. 

Diameter  of  Main  Crank-pin  Bearing  ............  4-J  in. 

Length  of  Main  Crank-pin  Bearing  ...............  5  in. 

Diameter  of  Front  and  Third  Parallel-rod  Bearings  3^  in. 
Length  of  Front  and  Third  Parallel-rod  Bearings  .  3^  in. 
Diameter  of  Second  Parallel-rod  Bearing  .........  5^  in. 

Length  of  Second  Parallel-rod  Bearing  ..........  4|  in. 

Cylinders  and  Steam-chests  ..................  Outside. 

Spread  of  Cylinders  ............................  84  in. 

Diameter  of  Cylinders  ..........................  19  in. 

Length  of  Stroke  ..............................  24  in. 

Position  of  Valve  Gear  ...............  Between  Frames. 

Type  of  Valve  Gear  ..............  Shifting-link  Motion. 

Travel  of  Valve  .................................  5  in. 

Outside  Lap  of  Valve  ............................  f  in. 

Inside  Lap  of  Valve  ............................  None. 

Lead  of  Valve  ..................................  ^  in. 

Throw  of  Eccentric  ..............................  5  in. 

Length  of  Steam  Ports  ........................  15|  in. 

Width  of  Steam  Ports  ..........................  1J  in. 

Width  of  Exhaust  Port  .........................  2|  in. 

Kind  of  Frames  .................  Wrought-iron,  Inside. 

Distance  between  Centers  of  Frames  ............  47  in. 

Boiler  Material  ................................  Steel. 

Thickness  of  Boiler  Sheets,  Dome  ...............  -fg  in. 

Thickness  of  Boiler  Sheets,  Barrel,  Outside  Fire-box 
and  Neck  .....................................  f  in. 

Smoke-box,  Waist,  and  Sheet  under  Dome  .......  iV  in- 


Maximum  Internal  Diameter  of  Boiler  } 


a 


53^  in. 


Minimum  Internal  Diameter  of  Boiler   i  51J  in. 

Height  to  Center  of  Boiler  from  Top  of  Rail  ......  76  in. 

Number  of  Tubes  ................................  119. 

Inside  Diameter  of  Tubes  .......................  2J  in. 

Outside  Diameter  of  Tubes  .....................  2|  in. 

Tube  Material  .........................  Wrought-iron. 

Length  of  Tubes  between  Tube  Sheets  ........  169  If  in. 

External  Heating  Surface  of  Tubes  .....  1,102.97  sq.  ft. 

Fire  Area  through  Tubes,  less  Ferrules  .....  3.29  sq.  ft. 


Length  of  Fire-box  at  Bottom  (Inside) 62£  in. 

Width  of  Fire-box  at  Bottom  (Inside) 34^  in. 

Height  of  Crown  Sheet  above  Top  of  Grate  (Center 

of  Fire-box) 53  in. 

Inside  Fire-box  Material Steel. 

Thickness  of  Inside  Fire-box  Sheets,  Sides J  in. 

Thickness  of  Inside  Fire-box  Sheets,  Front,  Back,  and 

Crown A  in. 

Thickness  of  Tube  Sheets ^  in. 

Tube  Sheet  Material Steel. 

Heating  Surface  of  Fire-box 92.96  sq.  ft. 

Total  Heating  Surface 1,195.93  sq.  ft. 

Fire-Grate  Area 15  sq.  ft. 

Maximum  Diameter  of  Smoke-stack  ) 

,,.   .  ,    i   Straight  . .  18  in. 

Minimum  Diameter  of  Smoke-stack  ) 

Height  of  Stack  from  Top  of  Kail 14  ft.  0  in. 

Width  of  Cab  Roof 9  ft. 

Height  of  Cab  Eoof  from  Rail 11  ft.  9f  in. 

Width  of  Cab  8  ft.  6  in. 

Size  of  Exhaust  Nozzle 3  x  4  in. 

Pressure  of  Steam  per  Square  Inch 125  Ibs. 

Nature  of  Fuel Bituminous  Coal. 

Weight  of  Engine,  Empty 77,000  Ibs. 

Weight  of  Engine  in  Working  Order 87,500  Ibs. 

Weight  on  First  Pair  of  Drivers 33,400  Ibs. 

Weight  on  Second  Pair  of  Drivers 29,500  Ibs. 

Weight  on  Third  Pair  of  Drivers 24,600  Ibs. 

Engine  fitted  with  Driver  Brake. 

Capacity  of  Tank 2200  gal. 

Capacity  of  Coal-box 3200  Ibs. 

Number  of  Wheels  under  Tender 8. 

Diameter  of  Wheels  under  Tender 33  in. 

Material  of  Wheels  under  Tender-trucks, 

Cast-iron  Chilled  Tread. 

Diameter  of  Tender-truck  Journals 3J  in. 

Length  of  Tender-truck  Journals 7  in. 

Weight  of  Tender,  Empty 21,300  Ibs. 

Weight  of  Tender,  Loaded 42,200  Ibs. 


CHAPTER    XVI. 

USEFUL  RULES,  FORMULAS,  AND   DATA. 
DIAMETER  OF  CYLINDERS. 

594.  In  Art.  23  we  have  given  one  method  of  computing  the  diameter  of  cylinders, 
when  the  stroke  of  piston,  mean  effective  steam  pressure,  and  weight  on  drivers  are 
known.  This  method  was  mainly  given  for  the  purpose  of  showing  the  principles 
upon  which  Rule  3  for  finding  the  tractive  force  was  based.  It  is  generally  best  to 
state  rules  of  this  kind  in  symbols ;  this  will  often  enable  us  to  very  readily  deduce 
from  them  other  rules.  Thus,  for  instance:  Let  D  denote  the  diameter  of  the 
driving  wheels  in  inches ;  d,  the  diameter  of  the  cylinder  in  inches ;  s,  the  stroke  of 
piston  in  inches ;  p,  the  mean  effective  steam  pressure  per  square  inch  of  piston ;  and 
T  the  tractive  force  in  pounds ;  then  Eule  3  will  take  the  following  form : 

d2  x  p  x  s 

— zr—  =  T-  w 

This  formula  gives  of  course  the  tractive  force,  and  this  multiplied  by  5  (see 
Art.  16)  gives  the  weight  on  drivers. 

From  formula  (a)  we  can  easily  deduce  a  simpler  rule  for  finding  the  diameter  of 
the  cylinder  than  is  given  in  Art.  23,  the  mean  effective  steam  pressure  on  the  piston, 
diameter  of  the  driving  wheels,  stroke  of  piston,  and  tractive  force  being  given. 

Multiplying  both  sides  of  formula  (a)  by  D,  we  get 

d2  x  p  x  s  =  T  x  D, 
from  which  we  obtain 

T  x  D 

d2  =  -       — » 
p  x  s 

hence, 


*J* 

V  p 


. 

which  in  ordinary  language  reads : 

EULE  115. — Multiply  the  tractive  force  in  pounds  by  the  diameter  of  the  driving 
wheel  in  inches ;  call  this  product  A.  Multiply  the  mean  effective  steam  pressure  per 
square  inch  of  piston  by  the  stroke  in  inches ;  call  this  product  B.  Divide  product  A 
by  product  B,  and  find  the  square  root  of  the  quotient,  which  will  be  the  diameter  of 
the  cylinder  in  inches. 

EXAMPLE  153. — What  should  be  the  diameter  of  the  cylinders  for  an  eight-wheeled 
passenger  engine — stroke  of  piston,  24  inches ;  mean  effective  steam  pressure,  90  pounds 

595 


596  MODERN  LOCOMOTIVE   CONSTRUCTION. 

per  square  inch  of  piston ;  diameter  of  driving  wheels,  61  inches ;  weight  on  drivers, 
57,360  pounds! 

Taking,  as  before,  |  of  the  weight  on  the  drivers  for  the  adhesion,  and  consequently 

for  the  tractive  force,  we  have 

57360 
—g—  =  11472  pounds. 

Now,  substituting  for  the  symbols  in  formula  (b)  their  values,  or  working  accord- 
ing to  Rule  115,  we  have 

d  =  y     QQ  x*24    =  V329  =  18  inches 

for  the  diameter  of  the  cylinders. 

In  the  foregoing  rule  we  have  neglected  the  internal  friction  of  the  engine,  which 
if  taken  into  account  would  somewhat  increase  the  diameter  of  the  cylinders  as  found 
above.  But  since  the  tractive  force  depends  on  the  adhesion,  and  since  the  latter  is 
in  many  cases  to  some  degree  an  uncertain  quantity — for  we  have  already  stated  that 
the  adhesion  varies  from  ^  to  J  of  the  weight  on  the  driving  wheels,  depending  on  the 
condition  of  the  rails — we  may  leave  the  internal  friction  of  the  engine  out  of  consider- 
ation ;  of  course  care  must  be  taken  not  to  adopt  too  great  a  tractive  force,  for  this 
would  cause  the  engine  to  be  over-cylindered. 

The  foregoing  rule  is  frequently  used  by  locomotive  builders,  and  agrees  veiy 
closely  with  the  one  recommended  in  the  Report  of  the  Proceedings  of  the  Twentieth 
Annual  Convention  of  the  American  Railway  Master-Mechanics'  Association. 

THE  MEAN   EFFECTIVE   STEAM  PRESSURE  REQUIRED  TO   DO   A  GIVEN  AMOUNT   OF  WORK. 

595.  From  formula  (a),  Art.  594,  we  may  deduce  another  one,  namely,  to  find  the 

mean  effective  pressure,  thus : 

Tx  D  .. 

r  =  d^Ts  (c} 

in  which  the  symbols  denote  the  same  quantities  as  given  in  Art.  594. 

In  ordinary  language  this  formula  reads : 

RULE  116. — Multiply  the  tractive  force  in  pounds  by  the  diameter  of  the  driving 
wheels  in  inches,  and  divide  this  product  by  the  square  of  the  diameter  of  the 
cylinder  in  inches  into  the  stroke  in  inches ;  the  quotient  will  be  the  required  mean 
effective  pressure  per  square  inch  of  piston. 

EXAMPLE  154. — Find  the  required  mean  effective  pressure  per  square  inch  of 
piston  in  an  eight-wheeled  passenger  locomotive  with  cylinders  18  inches  in  diameter, 
24  inches  stroke;  diameter  of  driving  wheels,  61  inches;  and  weight  on  driving 
wheels,  57,360  pounds. 

The  tractive  force  will  be 

57360 

— r —  =  11472  pounds, 
o 

According  to  formula  (c)  we  have 

11472  x  61 
P  =    182  x  24     =  9°  Pounds' 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


597 


that  is  to  say,  the  required  mean  effective  pressure  per  square  inch  of  piston  will  be 
90  pounds. 

596.  The  foregoing  rule  enables  us  to  compute  the  mean  effective  pressure  required  to 
do  a  given  amount  of  work.    But  frequently  the  question  arises :  What  will  be  the  mean 
effective  steam  pressure  when  the  boiler  pressure  and  the  point  of  cut-off  are  known  ? 

To  compute  this  with  any  degree  of  accuracy  we  must  take  the  clearance  into 
account.  In  commencing  a  new  design  of  a  locomotive  it  is  often  impossible  to  de- 
termine this  clearance  accurately,  and  therefore  for  preliminary  calculations  it  is 
frequently  neglected  until  the  design  is  far  enough  advanced  to  make  accurate  com- 
putations. But  in  either  case  there  are  other  rules  of  comparatively  simple  character, 
and  definitions  involved  to  which  we  shall  first  direct  attention. 

597.  In  all  computations  relating  to  the  distribution  and  action  of  steam  in  the 
cylinder  we  must  use  the  absolute  steam  pressure.    The  absolute  steam  pressure  is  equal 
to  the  sum  of  the  pressure  indicated  by  the  steam  gauge  plus  the  atmospheric  pressure ; 
the  latter  is  generally  taken  at  14.7  pounds  per  square  inch ;  consequently,  to  find  the 
absolute  pressure  add  14.7  pounds  to  the  gauge  pressure.     If  the  steam  gauge  indicates 
120  pounds,  then  the  absolute  pressure  will  be  120  +  14.7  =  134.7  pounds. 


IDEAL  INDICATOE-CAED. 

598.  In  order  to  gain  a  clear  conception  of  the  following  rules  it  will  be  advan- 
tageous to  understand  the  meaning  of  the  lines  which  form  the  outline  of  an  ideal 
indicator-card  such  as  is  shown  in  Fig.  1005.     The  horizontal  line  Z  0  is  (Jailed  the 
line  of  perfect  vacuum,  or  the  zero  line ;  from  this  line  all  the  absolute  pressures,  repre- 
sented by  lines  drawn  perpendicular  to  it,  are  laid  off.    The  length  of  the  line  Z  0  also 
represents  the  stroke  of  the  piston. 

The  line  Z  C  represents  the  absolute  initial  steam  pressure.     By  initial  steam 
pressure  is  meant  the  pressure  in  the  cylinder  at  or  near  the  beginning  of  the  stroke. 

The  line  C  D  represents  the  distance  through  which  the  piston  travels  before 
steam  is  cut  off;  hence  the  point  D  repre- 
sents the  point  of  cut-off. 

The  curve  D  F  represents  the  expan- 
sion curve,  and,  as  will  be  seen  later  on,  this 
curve  is  usually  assumed  to  be  a  rectangular 
hyperbola. 

599.  A  card  as  shown  in  Fig.  1005  can 
only  be  obtained  under  the  conditions  that 
the  exhaust  opens  exactly  at  the  end  of  the 
forward  stroke ;  closes  exactly  at  the  end  of 
the  return  stroke;  that  the  ports  are  fully 
opened  and  closed  instantaneously,  so  as  to 

avoid  wire  drawing  of  the  steam ;  and  that  there  is  no  compression.  Of  course  such 
conditions  never  exist  in  any  engine,  and  therefore  a  card  of  this  kind  cannot  be 
obtained  from  an  engine.  But  the  card  as  here  shown  serves  well  for  an  introduction 
to  the  study  of  indicator-cards,  and  is  an  aid  to  the  clear  understanding  of  this  subject. 


10       12       14 

Fig.  1005. 


598  MODERN  LOCOMOTIVE   CONSTRUCTION. 

600.  We  have  already  seen  that  the  line  Z  C  represents  the  pressure  at  the  begin- 
ning of  the  stroke ;  and  in  like  manner  any  other  line  drawn  perpendicular  to  Z  0  and 
terminating  in  the  line  C  D  or  the  curve  D  F  will  represent  the  pressure  at  a  corre- 
sponding point  of  the  stroke.     To  illustrate :  Let  the  zero  line  Z  0  (Fig.  1005)  repre- 
sent a  stroke  of  24  inches ;  then  the  line  12  e  drawn  through  the  center  of  Z  0  will 
represent  the  pressure  at  half  stroke ;  and  in  like  manner  the  line  F  0  represents  the 
terminal  pressure,  that  is,  the  pressure  at  the  end  of  the  stroke.     The  length  of  these 
lines  measured  with  a  suitable  scale  will  give  the  pressure  in  pounds  at  different  points  of 
the  stroke ;  of  course  the  same  scale  must  be  used  for  all  the  lines  of  pressure  in  a  card. 

601.  For  the  sake  of  simplicity  it  is  customary  to  assume  that  the  expansion  of 
steam  follows  the  law  of  Mariotte,  which  is  also  frequently  called  Boyle's  law.     A  curve 
D  F  laid  out  according  to  this  law  will  be  a  rectangular  hyperbola. 

Boyle's  law  states  that  the  pressure  of  a  perfect  gas  at  a  constant  temperature 
varies  inversely  as  the  space  it  occupies. 
From  this  we  get, 

pressure  x  volume  =  a  constant  quantity. 

Or,  let  p  denote  the  pressure  in  pounds ;  v,  the  volume ;  and  c  a  constant  quantity ;  we 

have 

p  x  v  =  c.  (a) 

Now  the  volume  of  the  cylinder  or  any  part  of  it  is  equal  to  its  area  multiplied  by 
its  length,  or  the  length  of  that  part  of  the  cylinder  which  is  under  consideration. 
Since  the  area  of  the  cylinder  is  constant  throughout  its  length,  we  need  not  pay 
any  attention  to  the  area  when  an  indicator-card  is  under  consideration,  and  therefore 
we  may  assume  that  the  zero  line  Z  0  denotes  the  whole  volume  of  the  cylinder,  and 
any  part  of  the  line  Z  0  represents  the  volume  of  a  corresponding  part  of  the  cylinder. 
Looking  at  the  length  of  the  line  Z  0  or  any  part  of  it  in  this  light  will  greatly 
simplify  matters,  as  will  be  seen  in  the  following  example. 

EXAMPLE  155. — Let  the  stroke  of  the  piston  be  24  inches ;  absolute  steam  pressure, 
150  pounds ;  cut-off  at  6  inches  of  the  stroke.  Find  the  pressure  at  half  stroke,  and  also 
the  terminal  pressure. 

In  this  example  we  have  p  =  150  pounds ;  v  =  6  inches ;  and  c  =  150  x  6  =  900. 
This  constant  quantity  will  not  change  for  any  part  of  the  stroke ;  hence  formula  (a) 

may.  in  this  case,  be  written 

p  x  v  =  900. 

For  half-stroke  v  =  12  inches ;  hence  we  have 

p  x  12  =  900, 
and  therefore  the  pressure  at  half  stroke  will  be 

900 
p  =  -y^r  =  75  pounds. 

For  the  terminal  pressure,  that  is,  the  pressure  at  the  end  of  the  stroke,  we  have 
v  =  24  inches,  and  therefore  the  terminal  pressure  will  be 

p  =  -qj-  =  37.5  pounds. 


MODERX  LOCOM01IVE   CONSTRUCTION. 


599 


In  a  similar  way  we  find  by  formula  (a)  the  pressure  at  any  other  part  of  the 
stroke;  thus  in  the  above  example  the  pressure  at  8  inches  from  the  beginning  of 

the  stroke  will  be 

900 
p  =  -Q-  =  112.5  pounds ; 


and  at  10  inches  from  the  beginning  of  the  stroke  the  pressure  will  be 

900 
10 


900 
p  =  -T^T  =  90  pounds ; 


and  so  on  for  any  other  part  of  the  stroke. 

602.  If  we  now  lay  off  these  pressures  on  corresponding  ordinates — that  is,  lines 
drawn  from  and  perpendicular  to  Z  0 — and  join  their  upper  extremities,  we  obtain 
the  rectangular  hyperbola  D  F.  Thus :  To  lay  out  the  expansion  curve  to  suit  the  con- 
ditions given  in  Example  155,  we  draw  the  line  Z  0  and  make  it  to  any  convenient 
scale  24  inches  long ;  divide  this  line  into  any  number  of  equal  parts,  say  12,  as 
shown  in  Fig.  1005 ;  then  the  distance  between  any  two  successive  points  will  represent 
2  inches ;  we  therefore  mark  the  points  of  division  (commencing  from  the  beginning 
Z  of  the  stroke)  2,  4,  6,  8,  etc. ;  through  these  points  of  division  draw  the  lines,  or 
ordinates,  perpendicular  to  Z  0.  Now,  steam  being  cut  off  at  6  inches,  the  ordinate 
6  D  must  be  made  to  represent  150  pounds,  which  is  the  absolute  initial  pressure  given 
in  the  example.  To  do  so  we  adopt  any  scale  of  pressure ;  this  scale  may  be  the  same 
one  as  used  for  laying  off  the  length  of  Z  0,  or  we  may  adopt  a  different  scale ;  but 
whatever  scale  we  do  adopt  to  lay  off  one  ordinate  must  be  used  for  all  the  other 
ordinates  in  the  card.  Let  us  adopt  1  inch 
to  represent  50  pounds,  then  6  D  must 
be  made  3  inches  long  to  represent  150 
pounds;  we  have  already  found  that  at 
8  inches  from  the  beginning  of  the  stroke 
the  pressure,  according  to  formula  (a),  will 
be  112.5  pounds,  and  therefore  the  ordinate 
on  point  8  must  be  2£  inches  long;  at  10 
inches  of  the  stroke  the  pressure  will  be  90 
pounds,  hence  this  ordinate  must  be  1$, 

inches  long.     In  a  similar  way  we  find  the  Fi    1( 

pressure   for  all  the  other  points  of  the 

stroke,  and  lay  off  these  pi-essures  on  corresponding  ordinates;   the  curve  drawn 
through  the  upper  extremities  of  these  ordinates  will  be  the  rectangular  hyperbola. 

This  curve  can  also  be  found  by  a  graphical  method,  thus :  Through  the  end  of  the 
stroke  0  (Fig.  1006)  draw  0/ perpendicular  to  Z  0,  and  prolong  CD  to  meet  0/in  the 
point/  thus  completing  the  rectangle  C  Z  Of.  Through  the  point  of  cut-off  D  draw 
D  d  perpendicular  to  Z  0;  through  Z  and  /  draw  the  diagonal  Zf,  cutting  D  d  in  the 
point^ ;  through  the  latter  point  drawy^  F  parallel  to  Z  0,  cutting  0/in  F;  this  point 
will  be  one  point  in  the  curve,  and  0  F  will  be  the  terminal  pi-essure.  Take  any 
point,  as  a,  on  the  line  Cf,  and  draw  the  diagonal  Z  a,  cutting  D  d  in  the  point  a, ; 
through  this  point  draw  the  line  a.2  «3  parallel  to  Z  0,  and  through  a  draw  a  «3  parallel 
to  Z  C,  meeting  a2  a3  in  the  point  a3,  which  will  be  another  point  in  the  curve.  By 


600  MODERN  LOCOMOTIVE   CONSTRUCTION. 

taking  any  number  of  points,  such  as  fe,  c,  on  the  line  C  f,  at  any  distance  apart 
(they  need  not  be  equal  distances  apart),  we  find  in  a  manner  similar  to  the  above  the 
points  &3,  c3  in  the  curve.  A  line  drawn  through  all  the  points  thus  found  will  be 
the  hyperbola. 

KATIO   OF  EXPANSION — CLEARANCE  NEGLECTED. 

603.  The  ratio  of  expansion  is  the  number  of  times  the  volume  of  that  part  of  the 
cylinder  occupied  by  the  steam  up  to  the  point  of  cut-off  is  contained  in  the  whole 
volume  of  the  cylinder.  Hence,  when  the  clearance  is  not  taken  into  account,  we  have 

volume  of  cylinder  , . 

— — -  -  =  ratio  of  expansion, 

volume  to  point  of  cut-off 

or, 

area  of  cylinder  x  length  of  stroke  , .      „ 

— —  =  ratio  of  expansion, 

area  of  cylinder  x  distance  to  point  of  cut-off 

From  which  we  get 

length  of  stroke  , .      ,, 

-  =  ratio  of  expansion. 

distance  to  point  of  cut-off 

RULE  117. — Divide  the  length  of  stroke  by  the  distance  from  the  beginning  of 
stroke  to  the  point  of  cut-off ;  the  quotient  will  be  the  ratio  of  expansion. 
This  rule  expressed  in  symbols  is  as  follows : 

f- 

in  which  L  denotes  the  length  of  the  stroke;  I,  the  distance  from  the  beginning  of 
the  stroke  to  the  point  of  cut-off ;  and  r,  the  ratio  of  expansion.  Of  course  when  L  is 
taken  in  inches,  I  must  also  be  taken  in  inches ;  or  if  one  is  taken  in  feet,  the  other 
must  also  be  taken  in  feet.  Or  if  L  is  regarded  as  1  (unity),  then  I  must  be  taken  as  a 
fraction  of  1. 

EXAMPLE  156. — Stroke  of  piston  is  24  inches ;  steam  is  cut  off  at  6  inches.    Find 

the  ratio  of  expansion. 

24 
According  to  Rule  117  we  have  -g-  =  4,  which  is  the  ratio  of  expansion. 

If  we  regard  the  stroke  as  1  (unity),  then  in  this  example  the  cut-off  must  be 

n 

regarded  to  take  place  at  HT  =  J  of  the  stroke ;  hence  we  have 

l^i=lxA=4 

for  the  ratio  of  expansion,  which  gives  us  the  same  result  as  before. 

EXAMPLE  157. — Stroke  of  piston  is  24  inches ;  steam  is  cut  off  at  16  inches.  Find 
the  ratio  of  expansion. 

Here  we  have 

24 

TJ.  =  1.5  for  the  ratio  of  expansion. 

Or,  since  16  inches  is  §  of  the  stroke,  we  have 

l-T-f  =  lxf  =  1.5  for  the  ratio  of  expansion. 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


601 


TERMINAL  PRESSURE. 

604.  We  have  already  seen  that  the  terminal  pressure  can  be  found  by  the 
formula  p  x  v  =  c  (see  Art.  601).  To  this  we  may  now  add  the  following : 

RULE  118. — Divide  the  initial  steam  pressure  by  the  ratio  of  expansion;  the 
quotient  will  be  the  terminal  pressure. 

EXAMPLE  158. — The  absolute  initial  steam  pressure  is  150  pounds;  stroke  of 
piston,  24  inches ;  steam  is  cut  off  at  6  inches.  Find  the  terminal  pressure. 

24 
According  to  Rule  117  the  ratio  of  expansion  is        =4;  and  according  to  Rule 

118  the  terminal  pressure  is 

150 

-7-  =  37.5  pounds. 

This  rule  holds  good  only  for  a  card  as  shown  in  Fig.  1005,  in  which  C  D  is 
parallel  to  Z  0,  and  the  exhaust  opens  at  the  end  of  the  stroke. 


MEAN    PRESSURE. 


605.  We  are  now  in  a  position  to  compute  the  mean  pressure  in  the  cylinder  when 
the  distribution  of  steam  is  as  indicated  in  Fig.  1005,  and  in  which  clearance  is  not 
taken  into  account. 

For  this  computation  we  need  a  table  of  hyperbolic  logarithms  as  given  below. 


TABLE  84. 

HYPERBOLIC    LOGARITHMS. 


Number. 

Hyperbolic  Log. 

Number. 

Hyperbolic  Log. 

1.1 

.095 

3.6 

1.281 

1.2 

.182 

3.7 

.308 

1.3 

.262 

3.8 

.330 

1.4 

.336 

3.9 

«?8P 

1.5 

.405 

4 

.3RG 

1.6 

.470 

4.1 

.411 

1.7 

.531 

4.2 

1.435 

1.8 

.588 

4.3 

1.459 

1.9 

.642 

4.4 

1.482 

2 

.693 

4.5 

1.504 

2.1 

.742 

4.6 

1.526 

2.2 

.788 

4.7 

1.548 

2.3 

.833 

4.8 

.569 

2.4 

.875 

4.9 

.589 

2.5 

.916 

5 

.609 

2.6 

.955 

5.1 

.639 

2.7 

.993 

5.2 

.649 

2.8 

1.030 

5.3 

.668 

2.9 

1.065 

5.4 

.686 

3 

1.099 

5.5 

.705 

3.1 

1.131 

5.6 

.723 

3.2 

1.168 

5.7 

.740 

3.3 

1.196 

5.8 

1.758 

3.4 

1.224 

5.9 

1.775 

3.5 

1.253 

6 

1.792 

To  find  the  hyperbolic  logarithm  of  any  other  number  not  given  in  the  table, 
multiply  the  common  logarithm  of  the  number  by  2.302585  (many  engineers  simply 
use  2.3  for  a  multiplier) ;  the  product  will  be  the  hyperbolic  logarithm  of  the  number. 


602  MODERN  LOCOMOTIVE  CONSTRUCTION. 

Tables  of  common  logarithms  are  given  in  nearly  all  engineers'  pocket-books  ;  they 
may  also  be  found  in  many  school-books,  and  need  not  be  given  here. 

For  finding  the  mean  pressure  we  have  the  following  : 

RULE  119.  —  Add  1  to  the  hyperbolic  logarithm  of  the  ratio  of  expansion  ;  multiply 
this  sum  by  the  absolute  initial  steam  pressure,  and  divide  the  product  by  the  ratio  of 
expansion  ;  the  quotient  will  be  the  absolute  mean  pressure  'per  square  inch. 

Or  in  symbols,  we  have 

1  +  hyperbolic  logarithm  r 

-  '     x  P  —  Pmi 

in  which  r  denotes  the  ratio  of  expansion  ;  p,  the  absolute  initial  steam  pressure,  and 
pm  the  absolute  mean  pressure. 

EXAMPLE  159.  —  The  absolute  initial  pressure  in  the  cylinder  is  180  pounds  ;  stroke 
of  piston,  24  inches  ;  steam  is  cut  off  at  6  inches  of  the  stroke.  Find  the  absolute 
mean  pressure. 

According  to  Eule  117  the  ratio  of  expansion  is 


The  hyperbolic  logarithm  of  4  (as  found  in  the  table)  is  1.386  ;  hence,  according 
to  Rule  119,  we  have 

x  180  =  107.37  for  the  absolute  mean  pressure. 


606.  So  far  we  have  not  taken  the  back  pressure  into  consideration  ;  if  we  assume 
it  to  be  uniform  throughout  the  stroke,  we  can  indicate  it  on  the  card,  Fig.  1005,  by 
drawing  the  line  P  p  parallel  to  Z  0  ;  the  distance  between  these  two  lines  must  be 
laid  off  with  the  same  scale  as  was  used  for  laying  off  the  initial  pressure.     If,  for 
instance,  we  have  used  a  scale  of  1  inch  to  represent  50  pounds,  and  the  absolute  back 
pressure  is  18  pounds,  then  the  distance  between  the  lines  Z  0  and  P  p  must  be  equal 
to  fo  =  -36  inch.     If,  now,  we  subtract  the  absolute  back  pressure  from  the  absolute 
mean  pressure,  the  remainder  will  be  the  mean  effective  pressure.     Thus:    In  the 
previous  example  we  have  found  the  mean  pressure  to  be  107.37  pounds  ;  if,  now,  the 
back  pressure  is  18  pounds,  the  mean  effective  pressure  will  be  107.37  —  18  =  89.37 
pounds. 

607.  Since  all  the  pressures  are  laid  off  from  the  zero  line  Z  0,  it  is  of  the  utmost 
importance  to  locate  it  correctly  on  a  card  taken   by  an   indicator.     To   do  so  the 
atmospheric  line  A  a  is  traced  by  the  pencil  on  the  card  before  it  is  taken  off  the 
indicator-drum.      When  this  line  is  to  be  traced  the  communications   between   the 
indicator  and  the  engine  cylinder  are  closed  ;  the  pencil  will  then  stand  in  its  neutral 
position,  and  will  trace  the  atmospheric  line  while  the  indicator-drum  revolves  on 
its  axis. 

When  the  card  is  taken  off,  the  zero  line  is  drawn  below  the  atmospheric  line,  and 
parallel  to  it  at  a  distance  which  will  represent  the  atmospheric  pressure  of  14.7 
pounds.  This  must,  of  course,  be  laid  off  with  the  adopted  scale  of  pressures  ;  if,  for 


MODERN  LOCOMOTIVE  CONSTRUCTION.  603 

instance,  a  scale  of  1  inch  to  represent  50  pounds  is  to  be  used,  the  distance  between 
the  atmospheric  and  zero  line  must  be  .294  inch. 

608.  It  should  be  distinctly  understood  that  any  scale  may  be  used  for  laying  off 
the  pressures  in  the  construction  of  an  ideal  card  such  as  is  shown  in  Fig.  1005,  but 
whatever  scale  we  may  adopt  must  be  used  for  laying  off  all  the  pressures  on  that 
particular  card.    Of  course,  when  the  pressures  are  to  be  measured  on  a  card  taken 
by  an  indicator,  we  must  use  a  scale  corresponding  to  the  spring  in  the  indicator, 
hence  in  cases  of  this  kind  there  is  no  choice  of  scale. 

609.  Under  the  assumption  that  the  expansion  of  steam  follows  Boyle's  law,  the 
mean  pressure  as  found  above  will  be  correct  when  there  is  no  clearance.    But  clear- 
ance we  have  in  all  engines,  because  the  steam  ports  have  volume,  however  short  they 
may  be  ;  and  besides,  some  play  between  the  piston  and  cylinder  heads  is  required  to 
allow  for  inaccurate  workmanship,  and  for  taking  up  the  wear  of  the  rod  brasses,  and 
that  between  the  driving  boxes  and  wedges.    It  must  also  be  remembered  that  in  loco- 
motives the  cut-off  is  regulated  by  the  link,  and  at  short  cut-offs  there  will  be  an  early 
exhaust  closure,  which  increases  the  compression  ;  and  in  order  to  prevent  the  pressure 
due  to  compression  becoming  too  great  we  require  probably  a  greater  clearance  than  in 
many  stationaiy  engines. 

The  clearance  space  being  filled  with  steam  at  the  moment  of  cut-off  will  have  an 
important  influence  on  the  ratio  of  expansion,  and  therefore  it  must  not  be  neglected  in 
accui-ate  computations.  Consequently  the  amount  of  clearance  must  be  determined 
accurately  ;  but  since  it  is  often  a  difficult  matter  to  compute  it  correctly,  a  practical 
method  is  frequently  adopted  by  placing  the  piston  at  the  end  of  the  stroke  and  fdling 
the  clearance  space  with  water  whose  quantity  in  cubic  inches  is  accurately  measured. 

610.  Our  next  step  will  be  to  indicate  the  clearance  on  the  card,  and  compute  its 
effect  on  the  expansion.     To  make  this  plain  we  will  take  the  following  example  : 

EXAMPLE  160.  —  The  cylinder  is  16  inches  in  diameter  ;  stroke  of  piston,  24  inches  ; 
cut-off  at  6  inches  ;  absolute  initial  steam  pressure,  180  pounds  ;  and  the  clearance  space 
contains  385  cubic  inches.  Find  the  ratio  of  expansion  and  the  mean  pi-essure  on  the 
piston. 

Our  first  step  will  be  to  find  the  ratio  between  the  clearance  and  piston  dis- 
placement. 

The  total  volume  of  the  piston  displacement  is  found  by  multiplying  the  area  of 
the  piston  by  its  stroke.  The  area  of  a  16-inch  piston  is  201  square  inches,  hence  the 
piston  displacement  is 

201  x  24  =  4824  cubic  inches. 

We  must  now  find  a  part  of  the  length  of  cylinder  which  will  contain  385  cubic  inches, 
the  clearance  space.  This  is  done  by  dividing  the  385  cubic  inches  by  the  area  of  the 
piston  ;  we  thus  obtain 

385 

=  1.91  inches, 


which  means  that  a  volume  16  inches  diameter  and  1.91  inches  long  contains  as  many 
cubic  inches  as  the  clearance. 

When  it  is  desirable  to  indicate  this  clearance  on  the  card  we  proceed  as  follows  :  Let 


604 


MODERN  LOCOMOTIVE  CONSTRUCTION, 


Z  to  0,  which  represents  the  stroke,  Fig.  1007,  be  4  inches  long,  then  the  distance  cor- 


responding to  1.91  inches  will  be 


1.91  x  4 
24 


=  .318  inch,  and  this  must  be  added  to  the 


card.  Hence  the  distance  from  Z  to  B,  which  represents  the  clearance,  must  be 
.318  inch.  Through  the  point  B  draw  a  line  perpendicular  to  Z  0,  and  complete  the 
rectangle  B  Z  E  C. 

If  there  had  been  no  clearance,  then  according  to  Eule  117  the  ratio  of  expansion 
would  have  been  4.  But  when  clearance  is  taken  into  account  the  ratio  of  expansion 
is  found  as  follows : 


or, 


Ratio  of  expansion  =        volume  of  cylinder  +  clearance^ 
volume  to  point  of  cut-off  +  clearance 


Ratio  of  expansion  = 


area  of  cylinder  x  (length  of  stroke  +  clearance) 


r  = 


area  of  cylinder  x  (length  of  stroke  to  point  of  cut-off  +  clearance) 

which  reduces  to 

length  of  stroke  +  clearance 

EULE  120. — Ratio  of  expansion  =  ,—  — r — T~  ~^  -~r~     —&- r  ~T~       — 

length  of  stroke  to  point  ot  cut-off  +  clearance 

To  express  this  rule  in  symbols,  we  have 

L  +  c 
I  +'c' 

in  which  r  denotes  the  ratio  of  expansion ;  L,  the  length  of  stroke  in  inches ;  /,  the 

distance  in  inches  from  the  beginning 
of  stroke  to  point  of  cut-off;  and  c,  the 
clearance. 

Substituting  for  the  symbols  the  val- 
ues given  in  the  example,  we  have 
24  +  1.91 


r  = 


=  3.2. 


8       10       12       14 

Fig.  1007. 


6  +  1.91 

From  this  it  will  be  seen  that  the  effect 
of  clearance  is  to  make  the  effective  cut- 
off comparatively  later  and  decrease  the 
ratio  of  expansion. 

In  the  foregoing  computation  the  values  of  L,  /,  and  c  must  always  be  expressed 
in  the  same  unit ;  if  one  is  taken  in  inches,  the  others  must  also  be  taken  in  inches ;  or 
if  one  is  expressed  in  feet,  the  others  must  also  be  expressed  in  feet. 

611.  Sometimes  the  clearance  is  given  in  per  cent,  of  the  piston  displacement.  In 
cases  of  this  kind,  L  =  1,  and  I  will  be  expressed  by  a  fraction  of  the  stroke  at  which 
steam  is  cut  off.  In  Example  160  the  clearance  is  eight  per  cent,  of  the  piston 
displacement ;  and  I  =  .25. 

Hence  the  ratio  of  expansion  is 

1  +  .08 
r  =  -^ ™  =  3.2,  as  before. 

.ZO   T"  .Uo 

For  finding  the  mean  pressure  on  the  piston  we  should  use  the  true  ratio  of 
expansion  as  found  above,  and  again  employ  Eule  119;  but  since  this  will  give  us 
the  mean  pressure  from  B  to  0  (Fig.  1007)  instead  of  from  Z  to  0,  the  length  of 


MODERN  LOCOMOTirE  CONSTRUCTION.  605 

the  stroke,  we  must  make  a  slight  correction,  as  will  presently  be  explained,  otherwise 
the  mean  pressure  will  be  too  high. 

We  have  found  that  with  clearance  the  ratio  of  expansion  is  3.2.  Now  the  hyper- 
bolic logarithm  of  3.2  is  1.163  ;  hence,  according  to  Rule  119,  the  mean  pressure  from 
B  to  0  (Fig.  1007)  will  be 

1      i      -|    -I  £»o 

x  180  =  121.66  pounds. 

O..J 

The  mean  pressure  from  Z  to  0  (Fig.  1007),  that  is,  the  length  of  the  stroke,  will  be 
less  than  found  above.  To  find  this  we  must  make  a  slight  correction  to  the  foregoing 
with  the  aid  of  the  following  rule  : 

RULE  121.  —  Multiply  the  mean  pressure  in  pounds  per  square  inch  as  found  by 
Rule  119  by  the  stroke  plus  the  clearance  in  inches;  multiply  the  initial  pressure 
by  the  clearance;  subtract  the  last  product  from  the  first  one;  and  divide  the  re- 
mainder by  the  stroke  in  inches  ;  the  quotient  will  be  the  mean  pressure  per  square 
inch  during  the  stroke. 

Or,  referring  to  Fig.  1007,  we  have  in  symbols, 

(Pm  x  BO)-  (p  x  BZ) 
ZO 

in  which  Pm  denotes  the  mean  pressure  during  the  stroke  ;  jpm,  the  mean  pressure  as 
found  by  Rule  119  ;  and  p,  the  initial  pressure. 

Substituting  for  the  symbols  their  values,  and  remembering  that  the  clearance 
measured  on  the  zero  line  is  1.91  inches,  we  have 


=  (121.66  x  25.91)  -  (180  xLglj  = 


24 

Hence  the  answer  to  example  is  :  Ratio  of  expansion,  3.2  ;  mean  pressure,  117  pounds. 
If,  now,  the  back  pressure  is  18  pounds,  the  mean  effective  pressure  will  be 

117  -  18  =  99  pounds. 

612.  To  sum  up  the  whole  matter,  we  have  the  following  mode  of  procedure  for 
finding  the  mean  effective  pressure  when  clearance  is  taken  into  account:  Find  the 
ratio  of  expansion  as  per  Rule  120  ;  find  the  mean  pressure  as  per  Rule  119  ;  correct 
this  mean  pressure  as  per  Rule  121,  and  subtract  the  back  pressure  ;  the  result  will  be 
the  mean  effective  pressure. 

HOESE-POWEB. 

613.  In  locomotive  practice  we  have   seldom  to  compute  the  horse-power;  Imt 
should  it  be  necessary  to  do  so,  we  must  find  the  mean  effective  pressure,  taking  into 
account  the  true  ratio  of  expansion  as  above,  and  when  this  is  known  the  horse-power 
is  veiy  easily  found  by  the  following: 

RULE  122.  —  Multiply  the  mean  effective  pressure  per  square  inch  of  piston  by  the 
area  of  the  piston  in  square  inches,  by  the  length  of  the  stroke  in  feet,  and  by  the 


606 


MODERN  LOCOMOTIVE  CONSTRUCTION. 


number  of  strokes  per  minute ;  divide  this  product  by  33,000 ;  the  quotient  will  be 
the  indicated  horse-power  for  one  cylinder. 

This  rule  may  be  expressed  in  symbols,  as  follows : 


/.  H.  P.  = 


P  x  L  x  A  x  N 

33000          ' 


in  which  P  denotes  the  mean  effective  pressure  per  square  inch;  L,  the  length  of 
stroke  in  feet  ;  A,  the  area  of  the  piston  in  square  inches  ;  and  N,  the  number  of 
strokes  per  minute. 

In  the  foregoing  formula  we  have  arranged  the  symbols  on  the  right-hand  side 
to  form  the  word  PLAN,  which  we  believe  will  be  an  aid  to  remember  the  rule.  It 
should  be  remarked  that  the  number  of  strokes  per  minute  is  equal  to  twice  the 
number  of  revolutions  of  the  wheels  in  the  same  time. 

EXAMPLE  161.  —  Diameter  of  cylinder  is  16  inches  ;  stroke,  24  inches  ;  cut-off  at  6 
inches  ;  absolute  initial  pressure,  180  pounds  ;  number  of  revolutions  of  wheels,  170 
per  minute  ;  clearance,  8  per  cent.  Find  the  indicated  horse-power. 

In  the  solution  of  Example  160  we  have  found  the  mean  effective  pressure  to  be 
99  pounds.  The  number  of  strokes  ler  minute  is  170  x  2  —  340,  and  the  area  of  a 
16-inch  piston  is  201  square  inches. 

According  to  Eule  122  we  have 


L 


99  x  2  x  201  x  340 
13000- 


=  41(X 


This  is  for  one  cylinder  ;  for  two  cylinders  we  have  410  x  2  =  820  L  H.  P. 
This  result  will  be  exact,  provided  the  mean  effective  pressure  P  has  been  correctly 
computed.  This  pressure,  as  found  in  the  last  example,  is  correct  for  a  card  as  shown 
in  Fig.  1007  ;  but  since  the  form  of  a  card  taken  from  an  engine  will  always  differ 
more  or  less  from  the  one  here  shown,  there  will  also  be  a  difference  in  the  mean 
effective  pressure  as  computed  above.  But  all  preliminary  computations  are  based 
upon  a  form  of  card  as  shown  in  Fig.  1007,  and  the  results  obtained  will  usually 

be  sufficiently  accurate  for  practical  pur- 
poses ;  or,  at  all  events,  they  will  be  a 
close  approximation. 

CARDS  TAKEN  WITH  AN  INDICATOR. 

614.  When  a  valve  gear  has  been  cor- 
rectly designed  we  may  expect  to  get, 
with  an  indicator,  a  card  similar  in  form 
to  the  one  shown  by  the  shaded  portion 
in  Fig.  1008,  when  steam  is  cut  off  at 
J  stroke.  The  distance  between  B  and  Z  represents  the  clearance,  which  has  been 
located  on  the  card  in  precisely  the  same  manner  as  that  adopted  for  locating  the 
clearance  line  in  Fig.  1007. 

It  will  be  noticed  that  in  this  card  the  point  of  cut-off,  exhaust  opening  and 


Fig.  1008. 


MODERN  LOCOMOTIVE  CONSTRUCTION.  607 

closure  is  not  as  decisively  indicated  as  in  the  ideal  cards  previously  shown,  and  this 
is  due  to  the  slowness  with  which  the  valve  opens  or  closes  the  ports.  Generally  the 
expansion  curve  from  D  to  G  does  not  coincide  exactly  with  the  rectangular  hyper- 
bola ;  and  from  G  to  F  a  decided  departure  takes  place. 

615.  In  this  diagram,  as  in  the  others,  the  line  A  a  is  the  atmospheric  line.    The 
line  Z  0,  the  zero  or  perfect  vacuum  line,  and  these  are  located  as  explained  in 
Art.  607. 

The  names  of  the  other  lines  are  as  follows : 

/  (7,   admission  line.  G  F,  exhaust  line. 

C  D,  steam  line.  F  //,  line  of  counter  pressure. 

D  G,  expansion  curve.  H  7,  compression  line. 

616.  From  a  diagram  of  this  kind,  taken  with  an  indicator,  the  mean  effective 
pressure,  and  consequently  the  horse-power,  when  required,  can  be  correctly  computed. 
The  mean  effective  pressure  is  usually  obtained  with  the  aid  of  the  planimeter,  which 
gives  the  area  of  the  card  (the  shaded  portion)  in  square  inches ;  dividing  this  area  by 
the  length  of  the  line  Z  0  (not  B  0)  measured  with  an  ordinary  rule,  we  obtain  the 
mean  height  of  the  card ;  and  this  multiplied  by  the  scale  of  the  spring  used  in  the 
indicator  will  give  us  the  mean  effective  pressure  per  square  inch  of  piston. 

The  mean  effective  pressure  can  be  obtained  from  this  card  in  another  way. 
Divide  the  length  of  the  card  Z  0  into  10  equal  parts,  thereby  obtaining  the  points  of 
division  d  d  d ;  through  the  center  of  each  division  draw  lines  perpendicular  to  Z  0 
or  A  a,  and  terminating  in  the  upper  lines  C  D  G  F.  Measure  the  lines  pr  rl5  p2  r2, 
etc.,  from  the  lower  boundary  line  I H  F  to  the  upper  one  C  D  G  F  with  an  ordinary 
rule.  Add  these  lengths  (all  in  inches),  and  divide  the  sum  by  the  number  of  lines ; 
the  quotient  will  be  the  mean  height  of  the  card.  Multiply  this  height  by  the  scale  of 
the  spring  used  in  the  indicator ;  the  product  will  be  the  mean  effective  pressure  per 
square  inch  of  piston. 

STEAM  ACCOUNTED  FOR  BY  THE  INDICATOR. 

617.  By  means  of  an  indicator-card  the  weight  of  steam  in  the  cylinder  at 
every  point  of  the  stroke  can  be  computed;   but  this  will  not  include  the  water 
existing  in  the  cylinder  at  the  time,  and  which  entered  as  steam ;  therefore,  the  steam 
accounted  for  by  the  indicator  is  not  the  full  weight  of  steam  that  entered  the 
cylinder,  or  the  full  weight  of  water  required  to  supply  the  boiler.    However,  by 
adding  a  certain  percentage  to  the  weight  of  steam  accounted  for  by  indicator,  we 
can  generally  determine  very  closely  the  amount  of  water  taken  out  of  the  boiler.     To 
make  this  plain  let  us  take  the  following  example : 

EXAMPLE  162. — Compute  the  weight  of  steam  accounted  for  by  the  indicator  from 
a  card  as  shown  in  Fig.  1008;  cylinder,  16  inches  diameter;  stroke,  24  inches; 
clearance,  8  per  cent.;  number  of  revolutions  per  minute,  170;  scale  of  spring,  80 
pounds  to  the  inch. 

Our  first  step  in  the  solution  of  this  problem  will  be  to  select  ;i  point  on  the 
expansion  curve,  as  near  as  possible  to  the  end  of  the  stroke,  at  which  it  is  certain  the 
exhaust  has  not  commenced.  The  card  shows  that  the  exhaust  commences  at  the 


608  MODERN  LOCOMOTIVE   CONSTRUCTION. 

point  Cr,  19  inches  from  the  beginning  of  the  stroke,  therefore  let  us  select  the  point 
L,  at  18  inches  from  the  beginning  of  the  stroke  ;  through  this  point  draw  a  line  L  M 
perpendicular  to  the  zero  line,  and  terminating  in  the  expansion  curve  at  one  end  and 
in  the  zero  line  at  the  other  end.  Now  measure  this  line  with  the  scale  of  the  spring  ; 
this  will  give  us  the  pressure  at  18  inches  from  the  beginning  of  the  stroke  ;  let  us 
assume  that  we  found  this  pressure  to  be  60  pounds  absolute. 

The  next  step  is  to  find  the  volume  occupied  by  the  steam  up  to  this  point  ;  this 
is  evidently  equal  to  sum  of  the  piston  displacement  up  to  this  point  plus  the  clearance 
space.  In  finding  this  volume  of  steam  we  must  remember  that  at  one  end  of  the 
cylinder  the  volume  is  reduced  by  the  space  occupied  by  the  piston-rod  ;  consequently 
we  must  take  the  mean  area  of  the  cylinder,  which  is  equal  to  the  area  corresponding 
to  the  diameter  of  the  cylinder  minus  one-half  of  the  piston-rod  area.  Let  us  assume 
that  the  latter  is  3  square  inches  ;  now  the  area  of  a  16-inch  cylinder  being  201  square 
inches,  we  have  for  the  mean  area  201  —  3  =  198  square  inches.  Consequently  the 
volume  occupied  by  the  steam  due  to  the  piston  displacement  alone  up  to  18  inches 
of  the  stroke  is  198  x  18  =  3564  cubic  inches.  To  this  we  must  now  add  the  volume 
due  to  the  clearance  space,  which  is,  as  given  in  the  example,  8  per  cent,  of  the  total 
piston  displacement  ;  and  the  latter  is  equal  to  the  mean  area  of  the  cylinder  multiplied 
by  the  whole  length  of  stroke,  hence  the  volume  due  to  the  clearance  alone  is  equal  to 
198  x  24  x  .08  =  380.16  cubic  inches.  Now  the  total  volume  of  steam  occupied  up 
to  18  inches  of  the  stroke  will  be  3564  +  380.16  =  3944.16  cubic  inches,  or 

3944.16 

=  2.282  cubic  feet. 


The  weight  of  a  cubic  foot  of  steam  at  a  pressure  of  60  pounds  (absolute)  is  .1457 
of  a  pound,  hence  the  total  weight  of  steam  in  the  cylinder  up  to  18  inches  of  the 
stroke  will  be  2.282  x  .1457  =  .3324  of  a  pound. 

From  this  we  deduct  the  weight  of  steam  saved  —  that  is,  the  steam  confined  in 
the  cylinder  after  the  exhaust  is  closed.  To  find  this  we  select  a  point  on  the  line  of 
counter-pressure  at  which  it  is  certain  that  the  exhaust  is  closed.  Let  us  take  the 
point  w,  4  inches  from  the  end  of  the  stroke.  The  volume  of  steam  confined  up  to  this 
point  will  be 

(198  x  4)  +  380.16 


1728 


=  .678  of  a  cubic  foot. 


The  absolute  pressure  at  the  point  n  is  found  by  measuring  its  distance  from  the 
zero  line  with  the  scale  of  the  spring ;  let  us  assume  that  in  this  way  we  find  the  ab- 
solute pressure  to  be  23  pounds  per  square  inch.  The  weight  of  a  cubic  foot  of  steam 
at  this  pressure  is  .0585  pounds,  hence  the  weight  of  steam  saved  is  .678  x  .0585  =  .0396 
of  a  pound. 

Now  the  quantity  of  steam  accounted  for  by  the  indicator  is  .3324  —  .0396  =  .2928 
pound  during  each  stroke.  The  wheels  make  170  turns  per  minute,  hence  the  number 
of  strokes  will  be  340  per  minute,  or  340  x  60  =  20400  strokes  per  hour,  and  the 
steam  used  in  one  cylinder  per  hour  as  accounted  for  by  the  indicator  will  be 

20400  x  .2928  =  5973.12  pounds. 


MODERN  LOCOMOTIVE   CONSTRUCTION.  609 

One  gallon  of  water  weighs  8.33  pounds,  hence  the  quantity  of  water  used  per  hour 

in  om>  cylinder  will  be 

5973.12 

"711  gallons, 


and  for  two  cylinders  the  amount  of  water  per  hour  as  accounted  for  by  the  indicator 
will  be  717  x  2  ==  1434  gallons. 

The  actual  amount  of  water  used  will,  of  course,  be  considerably  greater.  For 
late  cut-offs,  say  about  £  of  the  stroke,  we  should  add  20  per  cent,  for  cylinder  con- 
densation. For  early  cut-offs,  say  at  £  stroke,  we  should  add  35  per  cent,  for  cylinder 
condensation  ;  hence,  for  a  cut-off  at  £  stroke,  the  total  amount  of  water  used  will  be 

1434  x  1.35  =  1935.9  gallons. 

618.  From  the  foregoing  it  must  not  be  understood  that  a  greater  economy  is 
obtained  by  late  cut-offs;  indeed,  the  contrary  is  true.     But  the  above  does  show  that 
with  early  cut-offs  a  greater  cylinder  condensation   takes  place,  because  a  greater 
portion  of  the  cylinder  is  exposed  to  a  lower  temperature  at  the  end  of  each  stroke,  and 
therefore  a  greater  portion  of  the  cylinder  wall  is  also  cooled  ;  and  in  order  to  bring 
this  up  to  the  temperature  of  the  live  steam,  heat  must  be  expended,  and  consequently 
a  greater  condensation  will  take  place  than  when  a  lesser  portion  of  the  cylinder  is 
exposed  to  a  cooling  condition  as  occurs  in  late  cut-offs. 

619.  The  mean  effective  pressure  is  a  measure  of  the  work  done  in  the  cylinder  by 
the  steam  ;  hence  the  economy  in  the  use  of  steam  is  increased  by  making  the  mean 
effective  pressure  for  a  given  quantity  of  steam  as  high  as  possible  and  the  pressure  at 
the  opening  of  the  exhaust  as  low  as  possible,  thereby  obtaining  a  great  amount  of  work 
out  of  a  small  quantity  of  steam.     The  weight  of  steam  represents  the  weight  of  water 
that  must  be  evaporated  to  produce  it.    If,  now,  the  initial  pressure  is  150  pounds  per 
square  inch,  and  steam  is  cut  off  at  £  stroke,  the  mean  pressure  will  be,  according  to 
Rule  119, 

1  +  1.386 
—  —       x  150  =  89.4  pounds. 

If,  on  the  other  hand,  steam  follows  the  full  stroke  of  the  piston,  the  mean  pressure 
will  be  150  pounds  ;  but  in  this  case  we  use  four  times  as  much  steam  as  is  used 
in  cutting  off  at  J  stroke;  here,  then,  is  a  gain  with  an  early  cut-off,  but  some 
of  it  is  lost  by  the  greater  cylinder  condensation.  An  excellent  treatment  of  this 
subject  will  be  found  in  "  Indicator  Practice  and  Steam-engine  Economy,"  by  Frank 
F.  Hemenway  ;  this  book  also  contains  all  the  necessary  tables  for  making  computa- 
tions relating  to  the  use  of  steam. 

GRADE. 

620.  By  the  term  "  grade  "  is  meant  the  degree  of  inclination  of  the  track  from  a 
horizontal  line.     In  American  practice  the  grade  is  designated  by  the  rise  of  track  per 
mile;  for  instance,  if  it  is  said  that  the  grade  is  150  feet,  we  understand  that  iu  a 
length  of  one  mile  of  the  track  the  rise  is  150  feet, 


610 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


To  illustrate,  let  A  C  in  Fig.  1009  represent  one  mile  of  track ;  A  B,  a  horizontal 
line ;  and  B  (7,  a  line  drawn  vertically  to  A  B ;  then  B  C  will  represent  the  rise  of  the 
track  per  mile. 

According  to  some  text-books  treating  on  "  Mechanics, "  the  term  "  grade  "  would 
mean  the  rise  per  mile  of  the  horizontal  line  A  B ;  but  in  ordinary  railroad  practice  we 
generally  understand  it  to  mean  the  rise  per  mile  of  the  length  of  the  track  A  C;  this 
latter  definition  we  shall  adopt  in  that  which  is  to  follow. 

In  "  Mechanics  "  the  triangle  A  B  C  is  called  the  section  of  an  inclined  plane ;  A  C 
is  called  the  length  of  the  plane ;  A  B,  its  base ;  and  B  C,  its  height.  The  propositions 


_LL 


JJ_ 


Fig.  1010. 


1009. 


relating  to  a  body  moving  on  an  inclined  plane  as  given  in  "  Mechanics  "  are  also 
applicable  to  the  solutions  of  problems  of  a  train  moving  on  a  grade. 


NORMAL  PRESSURE. 

621.  In  the  footnote  on  page  6  we  have  stated  that  the  weight  and  pressure  are 
equal  only  on  a  level  or  horizontal  track,  and  not  on  a  grade;  we  shall  now  find 
the  difference  between  the  weight  and  pressure  of  a  body  on  an  inclined  plane. 

Let  TF,  Fig.  1010,  be  the  weight  of  a  body  resting  on  a  horizontal  plane  A  C;  each 
particle  of  this  body  gives  rise  to  a  reaction  which  is  equal  and  opposite  to  its  weight. 
We  may  therefore  say  that  the  total  reaction  of  this  body  is  equal  to  the  sum  of  the 
small  reactions,  which  may  be  replaced  by  their  resultant  R  passing  through  the  center 
of  gravity  of  the  body.  Consequently  this  resultant  E  is  equal  to  the  weight  of  the 
body,  but  acts  in  an  opposite  direction.  This  shows  us  that  the  normal  pressure  on  a 
horizontal  track  is  equal  to  the  weight  of  the  train. 

Let  Q,  Fig.  1009,  represent  a  body  resting  on  an  inclined  plane,  and  let  its  weight 
be  denoted  by  W.  Here,  again,  each  particle  of  the  body  causes  a  reaction  at  right 
angles  to  the  plane  A  C;  these  small  reactions  can  be  replaced  by  their  sum,  or,  in 
other  words,  by  their  resultant  E  acting  through  the  center  of  gravity  a  of  the  body 
perpendicular  to  A  C,  as  indicated  by  the  line  a  b,  and  this  resultant  is  equal  to  the 
normal  pressure  of  the  body  on  the  plane  A  C.  To  find  the  magnitude  of  this  result- 
ant we  apply  the  principle  of  the  parallelogram  of  forces,  thus : 

Through  the  center  of  gravity  a  of  the  body  draw  a  vertical  line  a  c  of  any  definite 


MOUKKX  LOCOMOTIVE   COXSTRUCTIOX.  611 

length,  to  represent  the  weight  W  of  the  body  Q.  Through  a  draw  a  line  a  b  perpen- 
dicular to  A  C,  and  through  c  draw  a  line  b  c  parallel  to  A  C,  meeting  a  b  in  b.  Now 
the  lino  a  c  represents  the  magnitude  of  the  force  or  weight  W,  and  the  line  a  b  repre- 
sents the  force  or  magnitude  of  the  resultant  E.  The  parallelogram  of  forces  teaches 
us  that  the  resultant  E  is  as  much  smaller  than  the  weight  W  as  the  line  a  b  is  shorter 
than  the  line  a  c.  Hence  we  have 

a  c  :  a  b  :  :  W  :  M,  (a) 

which  gives  us 

**^-£  =  A  (6* 

a  c 

Comparing  the  triangles  a  b  c  and  A  B  (7,  we  find  that  they  are  similar ;  hence 
formula  (a)  may  be  stated  as  follows : 

Length  of  inclined  plane  :  base  of  inclined  plane  :  :   W  :  E,          (c) 
which  gives  us  in  place  of  formula  (b) 


<-?  =  *,  W) 

AC 

base  of  inclined  plane  x  weight  of  body 

"f  inclined  plane  =  nomal  Pressure  R" 


Hence  to  find  the  normal  pressure  of  a  train  on  a  grade,  we  have  the  following  rule  : 
RULE  123.  —  Multiply  the  base  A  B  of  the  inclined  plane  by  the  weight  of  train,  and 

divide  this  product  by  the  length  A  C  of  the  inclined  plane  ;  the  quotient  will  be  the 

normal  pressure. 

EXAMPLE  163.  —  What  will  be  the  normal  pressure  of  a  train  weighing  120  tons  on 

a  grade  of  150  feet  per  mile  ?    From  the  definition  of  grade  given  in  Art.  620  we  know 

that  the  length  A  C,  in  this  case,  is  5,280  feet  (one  mile),  and  the  height  B  C  is  150  feet  ; 

the  length  of  A  B  must  be  found  by  computation,  which  need  not  be  given  here  ;  it  is 

sufficient  for  our  purpose  to  say  that  in  this  case  A  B  will  be  5279.4  feet. 
Now,  according  to  rule  we  have 

5279.4  x  120 

8  tons  normal  pressure. 


This  shows  us  that,  although  there  is  a  difference  between  the  normal  pressure 
and  the  weight,  it  is  so  small  that  it  may  be  neglected  ;  and  we  may  for  all  practical 
purposes  assume  that  the  normal  pressure  on  all  gi-ades  up  to  350  feet  per  mile  is 
equal  to  the  weight  of  train. 

RESISTANCE  DUE  TO   FEICTION. 

622.  In  hauling  a  train  up  an  inclined  plane  we  have  to  overcome  the  resistance 
due  to  rolling  and  axle  friction. 

When  the  surfaces  do  not  abrate  or  cut  one  another,  it  is  sufficiently  accurate  to 
assume  that  the  amount  of  friction  varies  directly  as  the  normal  pressure.  Hence, 
multiplying  the  normal  pressure  by  the  coefficient  of  friction,  we  obtain  the  amount  of 


612  MODERN  LOCOMOTIVE   COXSTRUCTION. 

friction.     In  Art.  621  we  have  shown  how  the  normal  pressure  on  an  inclined  plane  is 
found ;  if  we  now  denote  the  coefficient  of  friction  by  F,  we  have  (see  Fig.  1009) 

AB  x  W 


x  F  —  amount  of  friction. 


A  C 

We  have  already  seen  that  the  difference  between  the  normal  pressure  and  the 
weight  is  so  small  that  this  difference  can  be  neglected;  hence  in  ordinary  railroad 
practice  we  may  say  that 

Amount  of  friction  =  W  x  F. 

In  Art.  13  we  have  given  a  coefficient  of  friction  of  7£  pounds  per  ton,  and  in 
Art.  22  we  have  stated  that  this  coefficient  is  not  only  sufficient  to  overcome  the  resist- 
ance of  the  cars  due  to  friction,  but  it  is  also  sufficient  to  overcome  the  Motional 
resistance  of  the  engine.  Many  engineers  prefer  to  compute  the  resistances  of  cars 
and  engine  separately.  In  doing  so  a  coefficient  of  7  pounds  per  ton  is  generally 
adopted  to  overcome  the  rolling  and  axle  friction  of  the  cars,  the  same  is  also  adopted 
for  the  tender,  and  a  coefficient  of  8  pounds  per  ton  is  adopted  to  overcome  the  rolling 
and  axle  friction,  and  the  friction  of  the  machinery  of  the  engine.  These  coefficients 
are  only  suitable  when  the  train  is  in  motion.  For  starting  the  train  a  coefficient  of  8 
pounds  per  ton  should  be  used  for  overcoming  the  friction  of  the  cars,  and  10  pounds 
per  ton  for  the  engine. 

KESISTANCE  DUE   TO   GRADE. 

623.  In  hauling  a  train  up  a  grade,  work  is  performed.     By  the  term  "  work"  is 
meant  the  overcoming  of  resistances  continually  recurring  along  the  path  of  motion. 
To  do  work,  energy  must  be  expended.    By  the  term  "  energy  "  is  meant  the  ability  of 
an  agent  to  do  work.     That  which  gives  motion  to  the  body  is  force,  and  the  energy 
expended  is  equal  to  the  product  of  this  force  into  the  distance  through  which  it  acts. 
From  this  we  see  that  the  amount  of  energy  expended  is  expressed  in  foot-pounds,  or 
foot-tons,  etc.     Since  the  energy  expended  is  equal  to  the  work  done,  the  amount  of 
work  performed  is  also  expressed  in  foot-pounds,  or  foot-tons,  etc.     When  a  body  is 
moved  up  an  inclined  plane,   energy  must  be  expended  to  overcome  gravity  and 
friction.     Let  the  force  P,  Fig.  1009,  act  in   a  direction  parallel  to  the  plane  A  C; 
then  the  energy  expended  on  gravity  and  friction  in  moving  the  body  from  A  to  C 
is  equal  to  the  product  of  the  force  P  into  the  distance  A  C. 

624.  Let  us  determine  the  magnitude  of  the  force  P  required  to  overcome  the 
action  of  gravity  only,  the  corresponding  energy  expended,  and  woi'k  performed. 

This  force  P  is  found  by  the  principle  of  the  parallelogram  of  forces.  Hence  we 
construct  the  triangle  a  b  c,  in  Fig.  1009,  as  we  did  for  finding  the  reaction  E  in 
Art.  621.  Now  b  c  represents  the  magnitude  of  the  force  P,  which  is  as  much  smaller 
than  the  weight  W  of  the  body  Q  as  b  c  is  shorter  than  a  c.  But  the  triangles  a  b  c 
and  ABC  are  similar,  hence  we  have 


A  C  :  B  C  :  :  W  :  P,  (a) 

from  which  we  get 

7T77  x  W 

P  =  —          -•  (I) 

AC 


MODERN  LOCOMOTIVE  CONSTRUCTION.  613 

But  the  force  P  acts  through  the  distance  A  C,  hence  the  energy  expended  on 

gravity  is  found  us  follows : 

-B  n  x    ~uy        

Energy  expended  =  x  A  C  =  B  C  x  W.  (c) 

A  G 

But  the  work  performed  against  gravity  is  equal  to  the  energy  expended.  Hence 
we  may  say  the  work  done  on  gravity  is  equal  to  the  weight  W  raised  through  the 
height  B  C,  as  formula  (c)  indicates. 

ti'J.").  In  computing  the  hauling  capacity  of  a  locomotive  on  a  given  grade,  we  have 
to  deal  with  the  force  P;  formula  (b),  Art.  624,  shows  us  that 


B  C  x  W 
A~G 

In  this  formula,  B  C  is  the  rise  of  grade  in  feet  per  mile,  and  A  C  is  equal  to 
5,280  feet  (1  mile).     If  we  now  make  W  =  2000  pounds  (1  ton),  then  we  have 


p  _  B  C  x  2000 

5280 
in  which  B  C  must  be  taken  in  feet.    But 


0707 

~  "*'   '' 


5280 
hence  the  force  P  in  pounds  is 

P  =  B~C  x  .3787.  (d) 

From  this  we  get  the  following: 

RULE  124.  —  To  find  the  force  in  pounds  required  to  haul  a  train  of  one  ton  up  a 
given  grade,  under  the  assumption  that  there  is  no  friction:  Multiply  the  rise  in 
feet  per  mile  by  .3787  ;  the  product  will  be  the  force  required. 

RULE  125.  —  To  find  the  foi-ce  required  to  haul  a  train  of  any  weight  in  tons  up  a 
grade,  under  the  assumption  that  there  is  no  friction  :  Multiply  the  constant  .3787, 
as  given  in  Rulo  124,  by  the  weight  of  train  in  tons,  and  by  the  rise  of  grade  in 
feet  ;  the  product  will  be  the  required  force  in  pounds.  Of  course,  to  move  a  body  up 
an  inclined  plane  friction  cannot  be  neglected,  hence  the  force  P  must  not  only  be 
sufficiently  great  to  overcome  the  action  of  gravity,  but  it  must  be  able  to  overcome 
that  of  friction  also.  The  resistance  due  to  friction  has  been  given  in  Art.  622. 

In  order  to  show  the  application  of  the  foregoing  rules  for  finding  the  resistance 
of  a  train,  including  friction,  running  on  a  grade,  we  shall  take  the  following  example: 

EXAMPLE  164.  —  What  will  be  the  force  required  to  move  a  train  of  cars  and  tender 
up  a  grade  of  150  feet,  weight  of  train  100  tons? 
According  to  Art.  622,  we  have 

Resistance  due  to  friction  of  one  ton  (2,000)  =    7     pounds 

According  to  Rule  124, 

Resistance  due  to  grade,  150  x  .3787  =  56.8  pounds 

Total  resistance  per  ton  '  =  63.8  pounds 

The  total  resistance  due  to  the  weight  of  train  of  100  tons  will  be 

63.8  x  100  =  6380  pounds. 


614  MODERN  LOCOMOTIVE   CONSTRUCTION. 

That  is  to  say,  the  force  P  (Fig.  1009)  will  have  to  be  equal  to  6,380  pounds  to 
haul  a  train  of  the  given  weight  up  a  grade  of  150  feet,  at  a  very  low  rate  of  speed. 

RESISTANCE  DUE  TO   SPEED. 

626.  The  resistance  due  to  speed  depends  much  on  the  state  of  the  weather.    In 
calm  weather  the  resistance  is  less  than  in  strong  head-winds,  but  in  what  proportion 
the  resistance  increases   is  not  accurately  known ;    consequently,  for  computing  the 
resistance  due  to  speed  several  formulas  have  been  proposed,  which  give  results  at 
variance  with  each  other.     The  result  obtained  from  any  of  these  formulas  should  be 
considered  to  be  an  approximation  only.     The  following  will  give  probably  as  safe 
results  as  any  that  can  be  used  in  designing  engines  for  given  speeds : 

S2 

jYj-  -=  resistance  per  ton  (2,000  pounds)  of  train, 

in  which  S  denotes  the  speed  in  miles  per  hour.  This  rule  indicates  that  the  resist- 
ance increases  with  the  square  of  the  speed,  but  recent  observations  seem  to  indicate 
that  this  is  not  exactly  true,  and  the  resistance  is  less;  however,  the  error  resulting 
from  the  use  of  this  formula  will  be  on  the  safe  side.  From  this  we  get  the  following : 

RULE  126. — For  rinding  the  resistance  due  to  speed:  Divide  the  square  of  the 
speed  in  miles  per  hour  by  the  constant  171 ;  the  quotient  will  be  the  resistance  due  to 
speed  per  ton  of  load. 

Multiplying  this  quotient  by  the  weight  of  load  in  tons  will  give  the  resistance 
of  the  total  weight  due  to  the  speed. 

EXAMPLE  165. — What  is  the  total  resistance  of  the  load  (including  cars  and  tender), 
the  aggregate  weight  being  100  tons,  running  up  a  grade  of  150  feet  per  mile  at  a  rate 
of  30  miles  per  hour ! 

Resistance  due  to  axle  and  rolling  friction  =    7.0    pounds  per  ton 

Resistance  due  to  grade  =  150  x  .3787  =  56.8    pounds  per  ton 

2 

Resistance  due  to  speed  =  -$\  =    5.26  pounds  per  ton 

Total  resistance  per  ton  =  69.06  pounds 

Resistance  for  the  whole  weight  of  cars  and  tender  =  69.06  x  100  —  6906  pounds. 

EESISTANCE  DUE  TO   CURVES. 

627.  Curves  are  frequently  designated  by  degrees,  although  the  simplest  way  to 
designate  a  curve  is  by  its  radius,  to  which  locomotive  draftsmen  must  reduce  it  when 
given  in  degrees.     A  curve  of  one  degree  is  usually  assumed  to  have  a  radius  of  5,730 

feet ;  a  two-degree  curve,  — ^—  =  2865  feet ;  a  three-degree  curve,  — ^—  =  1910  feet ; 

and  so  on.  Hence,  if  the  degree  of  the  curve  is  given,  its  radius  is  approximately 
found  by  dividing  5,730  by  the  number  of  degrees.  This  rule  is  coirect  enough  for 
ordinary  curves  over  500  feet  radius. 

The  resistance  due  to  curves  is  influenced  by  many  circumstances.  It  depends  on 
the  length  of  train,  whether  the  cars  are  empty  or  loaded,  for  an  empty  train  will  offer 


MODERN  LOCOMOTIVE  CONSTRUCTION.  615 

a  greater  resistance  than  a  loaded  one  of  the  same  weight  ;  by  the  distance  between 
tli.'  wheels  under  each  car,  diameter  of  wheels,  shape  of  tread;  by  the  width  of  track, 
its  condition,  the  degree  of  elevation  of  the  outer  rails,  and  kind  of  coupling.  With 
so  many  conditions  entering  into  the  solution  of  a  problem,  it  is  easy  to  be  seen  that 
we  must  depend  on  experimental  data;  but  very  extensive  experiments,  as  far  as  we 
know,  have  not  been  made,  and  the  results  of  the  computation  based  on  the  insufficient 
data  at  hand  must  be  considered  only  as  approximations,  as  they  are  more  or  less  liable 
to  error.  The  following  rule  is  frequently  adopted  by  builders  : 

RULE  127.  —  Multiply  the  number  of  degrees  of  curvature  by  .5  ;  the  product  will 
be  the  resistance  in  pounds  per  ton  of  train  due  to  the  curves  on  a  road. 

This  rule  indicates  that  J  pound  per  ton  of  train  is  allowed  for  each  degree  of 
curvature. 

HAULING   CAPACITY  OP  LOCOMOTIVE  ON  A  GRADE. 

628.  We  are  now  in  a  position  to  compute  approximately  the  hauling  capacity  of 
a  locomotive  up  a  given  grade  with  a  curved  or  straight  track.  In  fact,  this  will  now 
be  a  very  simple  matter  ;  all  that  needs  to  be  done  is  to  fill  out  the  following  specifica- 
tion of  items,  and  then  proceed  as  shown  in  the  example. 

Resistance  of  cars  and  tender  due  to  rolling  i  ,  .      ,      .    , 

f  =  7  pounds  x  weight  of  cars  and  tender  in  tons. 
and  axle  friction,  in  pounds  J 

Resistance  of  cars  and  tender  due  to  grade.  >  ,        ,      . 

5  =  rise  x  .3787  x  weight  of  cars  and  tender  in  tons. 
in  pounds  } 

Resistance  of  cars  and  tender  due  to  speed,  >  =  (apeed  in  miles  per  hour)8  x          w  Q£  carg  and  teudej.  ^  ^^ 
in  pounds  )  171 

Resistance  of  cars  and  tender  due  to  curves,  >  ,  .      ,     .    . 

<  =  degrees  of  curve  x  .5  x  weight  of  carg  and  tender  in  tons. 
in  pounds  } 

Resistance  of  engine  due  to  rolling  and  asie  )  .  ,       .        .       .     . 

>  =  8  pounds  x  weight  of  engine  in  tons. 
friction  and  that  of  machinery,  in  pounds  ) 

Resistance  of  engine  due  to  grade,  in  pounds  =  rise  x  .3787  x  weight  of  engine  in  tons. 

Resistance  of  engine  due  to  speed,  in  pounds  =  -   x  weight  of  engine  in  tons. 

Resistance  of  engine  due  to  curves,  in  pounds  =  degrees  of  curve  x  .5  x  weight  of  engine  in  tons. 

EXAMPLE  166.  —  What  load  can  an  eight-wheeled  passenger  engine  with  cylinders 
18  x  24  inches  haul  up  a  grade  of  150  feet  per  mile,  the  road  having  a  curve  of  716 
feet  radius  on  the  grade?  Total.  weight  of  engine,  48  tons;  weight  on  drivers,  64,000 
pounds  ;  speed,  30  miles  per  hour. 

Since  the  sum  of  the  weight  of  cars,  with  load  and  weight  of  tender,  is  to  be  found, 
we  will  designate  the  sum  of  these  weights  by  W. 

The  radius  of  the  curve  is  given  at  716  feet  ;  hence,  according  to  ^rt.  627,  the  curve 

5730 
is  one  of  w-  =  8  degrees. 


Resistance  of  cars  and  tender  due  to  rolling  and  axle  friction  =  7  x  W                                 =7.0  W  pounds. 

Resistance  of  cars  and  tender  due  to  grade  =  150  x  .3787  x  W  =  66.8  H'  pounds. 

Resistance   of  cars  and   tender  due  to  speed  =  -          -  x  JT  =5.2  FT'  pounds. 

171 

Resistance  of  cars  and  tender  duo  to  curves  =  8  x  .5  x  W  =    4.0  W  pounds. 

Total  resistance  of  unknown  weights  =73.0  W  pounds. 


616  MODERN  LOCOMOTIVE  CONSTRUCTION. 

Resistance  of  engine  due  to  rolling  and  axle  friction  and  that  of  machinery  =  8  x  48        =      384     pounds. 
Resistance  of  engine  due  to  grade  =  150  x  .3787  x  48  =    2726.4  pounds. 

Resistance  of  engine  due  to  speed  =  —         -    x  48  =      249.6  pounds. 

171 

Resistance  of  engine  due  to  curves  =  8  x  .5  x  48  =      192     pounds. 

Total  resistance  of  known  weights  =    3552     pounds. 

The  adhesion  of  driving  wheels,  and  therefore  the  tractive  force,  is  one-fifth  of  the 

weight  on  drivers ;  consequently  we  have  — •= —  =  12800  pounds  to  overcome  a  total 

o 

resistance  of  3552  pounds  +  73  W  pounds. 

Subtracting  the  resistance  of  known  weights  from  the  tractive  force,  we  have 
12800  —  3552  =  9248  pounds  to  overcome  a  resistance  of  73  W  pounds ;  consequently 
the  weight  of  cars,  with  load  and  weight  of  tender,  will  be 

9°48 
W  =  -~-  =  126.68  tons, 

that  is,  the  given  engine  can  haul  under  the  conditions  given  in  the  example  126.68 
tons,  pi-ovided  the  boiler  has  a  sufficient  steaming  capacity  to  do  this  amount  of  work. 
Since  the  weight  of  train  as  found  above  is  that  which  the  engine  can  keep  in 
uniform  motion,  it  may  seem  that  the  engine  cannot  start  this  train,  because  the 
rolling  and  axle  friction  is  greater  at  starting  than  when  the  train  is  in  motion.  But 
it  must  be  remembered  that  in  the  foregoing  computation  we  have  made  an  allow- 
ance for  a  speed  of  30  miles  per  hour;  but  at  starting  the  speed  is  zero,  and  the 
allowance  we  have  made  for  speed  will  generally  be  sufficient  for  overcoming  the  extra 
friction  at  starting.  Hence  we  may  say  that  this  engine  will  also  start  the  train  on  the 
given  grade. 


CHAPTER    XVII. 

COMPOUND  LOCOMOTIVES. 

629.  A  compound  locomotive  is  one  in  which  the  expansion  of  steam  is  divided 
into  separate  stages.     Although,  strictly  speaking,  triple  expansion  engines  or  quad- 
ruple expansion  engines  are  compounds,  yet  custom  has  sanctioned  the  application  of 
the  term  "  compound  engine "  to  one  in  which  the  expansion  of  steam  is  divided  into 
two  stages  only.     The  general  American  locomotive  practice  at  present  is  to  use  one 
cylinder  for  each  stage  of  expansion.     The  steam  is  admitted  as  near  as  possible  to  the 
boiler  pressure  into  one  cylinder,  called  the  high-pressure  cylinder,  and  is  there  expanded 
to  a  greater  or  lesser  extent,  and  is  then  allowed  to  escape  into  a  second  cylinder  called 
the  low-pressure  cylinder,  where  it  is  again  expanded,  and  then  discharged  into  the 
atmosphere  in  the  usual  way. 

CYLINDER  CONDENSATION  AND  RE-EVAPORATION. 

630.  The  amount  of  cylinder  condensation  and  re-evaporation  vaiies  with  the 
difference  between  the  initial  and  final  temperature  in  the  cylinder.     Assume  that  the 
absolute  initial  pressure  in  a  simple  engine  is  165  pounds ;  at  this  pressure  the  steam 
has  a  temperature  of  365.7  degrees  Fahr. ;  if  the  final  pressure  is  33  pounds  absolute, 
the  temperature  of  the  steam  will  then  be  255.7  degrees,  giving  a  difference  of  tempera- 
ture of  365.7  —  255.7  =  110  degrees,  in  the  cylinder  during  every  stroke.     Of  course  it 
cannot  be  claimed  that  the  range  of  temperature  in  the  metal  of  the  cylinder  walls  will 
be  as  great  as  the  difference  between  the  initial  and  final  temperature  of  the  steam ; 
but  that  there  is  a  difference  in  the  temperature  of  the  cylinder  walls  during  each  stroke 
is  beyond  a  doubt,  and  we  naturally  infer  from  this  that  when  the  entering  steam 
comes  in  contact  with  the  surface  of  the  metal  whose  temperature  is  below  that  of  the 
steam,  condensation  takes  place  and  reduces  the  work  which  otherwise  the  steam  could 
perform. 

631.  As  the  steam  is  expanded  its  temperature  falls  below  that  of  the  cylinder 
walls,  causing  heat  to  be  transferred  from  the  latter  to  the  former,  and  causing  a  re- 
evaporation  ;  and  when  the  exhaust  takes  place,  much  of  the  heat  transferred  from  the 
cylinder  walls  to  the  steani  is  lost,  causing  a  loss  of  work. 

632.  We  have  seen  that  in  a  compound  engine  the  expansion  is  divided  into  two 
stages,  each  taking  place  in  a  separate  cylinder.     This  reduces  the  variation  of  press- 
ure and  temperature  in  each  cylinder,  and  therefore  it  is  said  that  the  cylinder  con- 
densation in  a  compound  is  less  than  in  a  non-compound,  the  total  ratio  of  expansion 
being  the  same  in  both  engines.     But  the  advocates  of  simple  engines,  among  whom 

617 


618  MODERN  LOCOMOTIVE    CONSTRUCTION. 

there  are  many  engineers  whose  reputation  gives  weight  to  their  opinion,  point  to  the 
fact  that  in  a  compound  engine  the  steam  is  exposed  to  a  greater  cylinder  cooling  sur- 
face, and  therefore  insist  that  the  cylinder  condensation  is  not  reduced  as  much  as  is 
claimed  by  the  advocates  of  the  compound  engines,  and  that  the  saving  of  fuel  in  the 
latter  class,  if  any,  cannot  be  due  to  a  reduced  cylinder  condensation.  This  opinion 
cannot  well  be  refuted,  because  it  is  extremely  difficult  to  detect  the  cylinder  conden- 
sation on  the  indicator-cards  taken  when  the  locomotives  are  running  at  a  high  speed 
and  cutting  off  at  an  early  part  of  the  stroke. 

633.  In  a  compound  engine,  the  steam  due  to  re-evapoi'ation  that  takes  place  in 
the  high-pressure  cylinder  is  not  wholly  lost,  because  it  is  admitted  into  the  low-press- 
ure cylinder  and  does  work  there.     The  re-evaporation  in  the  low-pressui-e  cylinder, 
however,  causes  a  loss,  because  the  steam  due  to  it  will  be  discharged  into  the  atmos- 
phere before  it  has  a  chance  of  doing  a  corresponding  amount  of  work ;  but  this  loss 
will  not  be  as  great  as  in  a  simple  engine. 

634.  There  is  certainly  a  gain  in  using  steam  at  a  high  pressure  and  cutting  off  short 
so  as  to  expand  it  to  as  low  a  terminal  pressure  as  is  practical.     To  do  so  a  difficulty  is 
encountered.     The  ratio  of  expansion  is  governed  by  the  link  motion,  which  is  usually 
of  the  type  shown  in  Fig.  29,  and  on  account  of  its  simplicity  is  retained  in  compounds. 
With  this  gear  the  distribution  of  steam  in  each  cylinder  is  governed  by  one  valve,  and 
this  may  cause  considerable  difficulty  in  the  attempt  to  keep  the  pressure  due  to  com- 
pression at  early  cut-offs  within  desirable  limits;  and  indeed  this  matter  is  not  so 
easily  disposed  of  in  compounds  as  in  simple  engines.     The  clearance  space  affects  the 
pressure  due  to  compression.    For  smooth  running  it  should  be  sufficient  to  arrest  the 
motion  of  the  reciprocating  parts ;  this  is  a  problem  by  itself,  and  need  not  be  con- 
sidered here.     For  economy  of  steam  the  pressure  at  the  end  of  compression  should  be 
equal  to  the  initial  pressure ;  there  is  no  advantage  in  compressing  the  steam  to  a  higher 
degree.    Now,  if  in  a  simple  engine  the  absolute  initial  pressure  is  160  pounds  and  the 

-i  (*f\ 

absolute  back  pressure  is  18  pounds,  we  may  compress  the  latter  -TTT  =  8.8  times  before 

the  initial  pressure  is  reached.  If  in  a  compound  the  initial  pressure  in  the  high-pressure 
cylinder  is  160  pounds  and  the  back  pressure  in  the  low-pressure  cylinder  is  18  pounds, 
we  have  the  same  range  of  compression,  but  this  takes  place  in  two  cylinders ;  if,  for 
instance,  the  receiver  pressure  is  65  pounds,  then  the  range  of  compression  in  the  H.  P. 
cylinder  will  be  from  65  to  160  pounds,  and  in  order  to  reach  the  initial  pressure  the 

1  CC\ 

steam  will  have  to  be  compressed  about  -^~  =  2.4  times.     In  the  L.  P.  cylinder  the 

range  of  compression  is  from  18  to  65  pounds,  and  therefore  the  back  pressure  will 

65 

have  to  be  compressed  about  ^  =  3.6  times  before  the  receiver  pressure  is  reached. 

lo 

Now  if  the  same  kind  of  link  motion  and  the  same  size  of  valves  are  used  in  both 
classes  of  engines,  it  will  easily  be  perceived  that  there  is  danger  of  obtaining  an 
excessive  and  hurtful  pressure  due  to  compression,  unless  the  clearance  space  is  larger 
in  the  compound  than  in  the  simple  engine.  From  this  it  will  be  seen  that  a  large 
percentage  of  clearance  space  may  be  of  great  advantage  in  compound  engines.  There 
is  another  way  of  avoiding  an  excessive  compression,  namely,  giving  the  valve  a  large 


MODERN  LOCOMOTIVE   CONSTRUCTION.  619 

inside  clearance  so  as  to  secure  a  late  exhaust  closure  ;  this  is  often  resorted  to,  and  in 
fact  it  is  the  noticeable  difference  between  the  slide  valves  of  a  compound  and  an 
ordinary  locomotive.  Hence,  an  increased  percentage  of  clearance  space,  or  inside 
clearance  of  the  valves,  or  both,  may  be  advantageously  employed  in  compound  locomo- 
tives. 

635.  The  point  at  which  the  exhaust  closure  should  take  place  so  as  to  give  any 
desired  pressure  at  the  end  of  compression  can  be  found  by  the  following  formula  : 


in  which  d  denotes  the  distance  H  F  (see  Fig.  1008),  or,  in  other  words,  the  distance 
from  the  beginning  F  of  the  return  stroke  to  the  point  H  of  the  exhaust  closure  ;  P, 
the  desired  absolute  pressure  B  E  at  the  end  of  compression,  which  may  be  equal  to 
the  initial  pressure  or  less  ;  jp,  the  absolute  back  pressure  ;  and  c,  the  clearance  B  Z 
in  per  cent,  of  the  stroke. 

In  ordinaiy  language  the  above  formula  reads  : 

RULE  128.  —  Divide  the  desired  absolute  pressure  at  the  end  of  compression  by  the 
absolute  back  pressure  ;  subtract  1  (one)  from  the  quotient,  and  multiply  the  remainder 
by  the  clearance  in  per  cent,  of  the  stroke  ;  subtract  this  product  from  1  (one)  ;  the 
remainder  will  be  the  distance  in  per  cent,  from  the  beginning  of  the  return  stroke  to 
the  point  at  which  the  exhaust  valve  should  close. 

EXAMPLE  167.  —  The  clearance  space  is  8  per  cent.,  the  pressure  at  the  end  of 
compression  is  to  be  160  pounds  absolute,  and  the  absolute  back  pressure  is  18 
pounds  :  at  what  distance  from  the  beginning  of  the  return  stroke  should  the  exhaust 
valve  close! 

Substituting  for  the  symbols  in  the  above  formula  the  values  given  in  the 
example,  we  have 

d=  1  ~(^-    -  l)x  .08  =  .37, 

that  is  to  say,  the  exhaust  should  close  at  .37th  part  of  the  stroke  ;  if  the  latter  is  24 
inches,  then  the  exhaust  should  close  at  24  x  .37  =  8.88  inches  from  the  beginning  of 
the  return  stroke. 

636.  With  the  same  valves  and  valve  gear  in  both  classes  of  engines,  we  can 
always  obtain  a  greater  ratio  of  expansion  in  a  compound  than  in  a  simple  engine, 
because  in  cutting  off  at  the  same  point  in  both  classes  the  steam  in  the  high-pressure 
cylinder  will  be  expanded  to  the  same  extent  as  in  the  simple  engine;  but  this  is  again 
expanded  in  the    low-pressure  cylinder,   thereby  obtaining    an   increased  ratio  of 
expansion,  and  therefore  high-pressure  steam  can  be  used  with  better  advantage  in 
a  compound  than  in  an  ordinary  locomotive.    Again,  it  must  not  be  overlooked  that 
the  expansive  working  of  steam  is  compulsory  in  a  compound;  even  with  the  link 
motion  set  in  full  gear,  there  will  be  an  expansion  of  steam  in  the  compound  greater 
than  in  a  simple  engine,  and  this  contributes  much  to  the  success  of  the  former. 

637.  In  compounds  the  variation  of  steam  pressure  during  a  stroke  is  reduced, 
consequently  the  pressures  on  tho  crank-pins  are  more  uniform,  which  reduces  the 
tendency  to  slip  the  wheels,  and  gives  greater  durability  to  the  parts. 


620  MODERN  LOCOMOTIVE   CONSTRUCTION. 

638.  The  lack  of  reliable  experiments  leaves  the  amount  of  fuel  that  can  be  saved 
with  a  compound  in  dispute ;  indeed,  the  reports  relating  to  this  are  very  conflicting. 
Considerable  of  this  dispute  is  probably  due  to  the  prejudice  of  the  advocates  of 
the  simple  engine  on  the  one  hand,  and  the  enthusiasm  of  the  advocates  of  the  com- 
pound on  the  other  hand,  who  are  liable,  though  perfectly  honest,  to  overestimate  the 
economy  of  fuel.     Eeports  seem  to  indicate  that  in  a  very  few  cases  a  good  simple 
engine  has  given  as  good  results  as  a  compound ;  but  in  the  majority  of  cases  the 
compounds  have  given  much  better  results.     It  is  probably  safe  to  say  that  the  saving 
of  fuel  in  favor  of  the  compound  will  vary  from  15  to  30  per  cent. ;  the  latter  amount 
may  be  expected  when  the  compound  is  doing  the  work  for  which  she  has  been 
designed,  and  this  saving  will  be  decreased  when  the  engine  has  more  or  less  than  this 
amount  of  work  to  do. 

TWO-CYLINDER  COMPOUNDS. 

639.  The  two-cylinder  compounds  may  be  divided  into  two  distinct  classes.     First, 
the  class  which  embraces  those  which  always  work  as  compounds  except  for  a  brief 
period  at  starting ;  they  are  so  arranged  that  they  change  automatically  from  simple 
to  compound  working.     Second,  the  class  which  embraces  those  which  can  be  worked 
as  simple  or  as  compounds  at  the  will  of  the  engineer. 

The  outer  appearance  of  either  class  does  not  differ  much  from  that  of  the  simple 
engine ;  indeed,  it  does  not  require  many  changes  to  convert  a  simple  into  a  compound 
engine. 

LOCOMOTIVES  WHICH  ALWAYS  WORK  AS  COMPOUNDS. 

640.  Fig.  1013  shows  a  partial  end  view  of  an  engine  which  always  works  as  a 
compound,  excepting  during  the  brief  period  at  starting.     The  high-pressure  cylinder 
is  marked  H.  P.,  and  the  low-pressure  cylinder  L.  P. ;  the  latter  is  of  course  larger  in 
diameter  than  the  former.     An  ordinary  steam  pipe,  not  shown  here,  but  similar  in 
design  to  that  shown  in  Fig.  24,  is  used  to  convey  the  steam  from  the  boiler  into  the 
H.  P.  cylinder ;  after  it  has  done  work  there  it  is  discharged  into  the  pipe  r,  which 
conveys  the  steam  into  the  L.  P.  cylinder ;  this  pipe  r  is  called  the  receiver ;  after  the 
steam  has  done  its  work  in  the  L.  P.  cylinder  it  is  discharged  into  the  atmosphere  in 
the  usual  way.     The  steam  in  its  passage  to  the  L.  P.  cylinder  has  to  pass  through  the 
intercepting  valve  placed  in  the  saddle  as  shown,  and  of  which  different  sections  are 
given.     This  valve  is  the  invention  of  Mr.  Albert  J.  Pitkin,  superintendent  of  the 
Schenectady  Locomotive  Works. 

Fig.  1011  is  a  longitudinal  elevation  partly  in  section  of  the  L.  P.  cylinder ;  the 
partial  section  is  taken  on  the  line  5  5  of  Fig.  1013. 

Fig.  1012  is  a  longitudinal  sectional  plan  of  the  saddle  castings  of  the  H.  P.  and 
L.  P.  cylinders,  and  the  regulator  mechanism  taken  on  the  line  1  1  of  Fig.  1013. 

Fig.  1014  is  a  longitudinal  sectional  elevation  of  a  portion  of  the  regulator,  taken 
on  the  line  2  2,  Fig.  1012,  and  line  3  3,  Fig.  1013. 

Fig.  1015  is  a  transverse  sectional  elevation  of  the  regulator,  taken  on  the  line 
4  4,  Fig.  1012. 


Fig.  1011. 


622 


MODERN  LOCOMOTIVE   CONSTRUCTION. 


MODERN  LOCOMOTIVE  CONSTRUCTION.  623 

Fig.  101(5  is  a  transverse  sectional  elevation  of  the  regulator,  taken  on  the  line 
6  6,  Fig.  1012. 

Fig.  1017  is  a  longitudinal  elevation  of  the  H.  P.  cylinder. 

Fig.  1018  is  a  plan  of  the  saddle  connections  of  the  H.  P.  cylinder. 

Fig.  1019  is  a  longitudinal  sectional  elevation  of  a  portion  of  the  regulator  on 
the  line  7  7,  Fig.  1012. 

Fig.  1020  is  a  transverse  sectional  elevation  of  the  intercepting  valve  and  chamber 
on  the  line  8  8,  Fig.  1011. 

Fig.  1021  is  a  longitudinal  sectional  plan  of  the  regulator  and  intercepting  valve- 
chamber,  taken  on  the  line  1  1,  Fig.  1013,  showing  the  several  parts  in  opposite 
positions  to  that  in  Fig.  1012. 

The  action  of  these  valves  is  described  as  follows :  Assuming  the  engine  to  be  at 
rest,  upon  opening  the  throttle  live  steam  from  the  boiler  enters  the  H.  P.  cylinder 
steam-chest  through  the  passage  c,  Figs.  1012  and  1018,  while  simultaneously  this  live 
steam  passes  through  the  pipe  p  into  the  passage  p'  of  the  regulator.  From  this  pas- 
sage it  enters  the  auxiliary  valve-chamber  /  and  forces  the  auxiliary  valve  v"  into  the 
position  shown  in  Fig.  1014.  The  live  steam  then  flows  through  the  port y,  Fig.  1014, 
into  the  valve-chamber  «,  Fig.  1012.  Owing  to  the  difference  in  areas  of  the  pistons 
t  and  t'  the  preponderance  of  the  pressure  will  cause  the  pistons  t  and  V  with  the 
valve  v1  to  move  to  the  opposite  end  of  the  chamber  n  from  that  shown  in  Fig.  1012. 
This  movement  of  the  valve  uncovers  the  port  x  and  allows  the  live  steam  to  enter  the 
cylinder  h.  This  pressure  upon  the  piston  /  will  cause  it  to  move  to  the  opposite  end 
of  the  cylinder  h  from  that  shown  in  Fig.  1012.  This  movement  of  the  piston  I 
through  the  rod  I'  causes  the  intercepting  valve  d  to  close,  thus  cutting  off  all  commu- 
nication between  the  high-  and  low-pressure  cylinders,  as  shown  in  Fig.  1021. 

The  angularity  of  the  two  arms  of  the  lever  g  in  relation  to  the  fixed  position  of 
the  collar  <f  upon  the  piston-rod  I'  is  such  that,  when  the  intercepting  valve  is  being 
moved  forward  and  closing,  the  shorter  arm  of  the  lever  g  will  strike  the  end  of  the 
spindle  s'  at  the  time  the  intercepting  valve  d  has  just  closed  line  and  line  with  its 
seat.  The  further  movement  of  the  intercepting  valve,  owing  to  its  lap,  will  be  suffi- 
cient to  cause  the  valve  v  to  acquire  its  full  opening.  As  soon  as  the  valve  v  is  opened 
the  live  steam  from  the  passage  JM'  flows  through  the  opening  o,  the  clmniher  m,  the 
opening  made  by  the  valve  v,  the  passage  p",  through  the  intercepting  valve  chamber 
d'  and  the  opening  o'  into  the  receiver,  and  thence  on  to  the  steam-chest  of  the  low- 
pressure  cylinder.  This  pressure  being  upon  the  under  side  of  the  intercepting  valve, 
should  any  lost  motion  exist,  raises  it  against  the  upper  side  of  its  chamber,  and  pre- 
vents any  live  steam  from  passing  back  through  the  receiver  and  acting  as  back  press- 
ure on  the  high-pressure  cylinder.  Stated  concisely,  the  operation  of  starting  is  tlius: 
Upon  opening  the  throttle,  live  steam  is  admitted  simultaneously  to  both  the  high- 
and  low-pressure  cylinders,  and  by  means  of  this  same  live  steam  acting  through  a 
mechanism,  separate  and  distinct  from  the  intercepting  valve  itself,  the  intercepting 
valve  is  closed  automatically  and  the  engine  starts  with  its  full  ]><>\vcr  as  simple  or  non- 
compound  engine.  The  engine  now  starting,  the  exhaust  steain  from  the  II.  P.  cylinder 
is  accumulating  in  the  receiver  above  the  intercepting  valve.  This  increasing  pressure 
in  the  receiver  is  carried  through  the  pipe  p"'  into  the  auxiliary  valve  chamber  f,  Fig. 


624  MODE  JIN  LOCOMOTIVE   CONSTRUCTION. 

1014,  thence  through  the  ports  x'"  into  the  valve  chamber  «,  Fig.  1012,  and  exerts  itself 
on  the  side  of  the  piston  t',  Fig.  1012,  opposite  to  that  acted  upon  by  the  live  steam. 
As  soon  as  the  requisite  pressure  has  accumulated  in  the  receiver,  and  which  is  sufficient 
to  overcome  the  live  steam  pressure  on  the  unequal  inner  areas  of  the  pistons  t  and  t', 
these  pistons  with  the  slide  valve  v'  are  caused  to  assume  their  original  position,  as 
shown  in  Fig.  1012,  thus  uncovering  the  port  x',  allowing  the  live  steam  from  the 
chamber  n  to  enter  the  corresponding  end  of  the  cylinder  h.     At  the  same  time  the 
slide  valve  v'  prevents  the  further  entrance  of  the  live  steam  to  the  cylinder  h  through 
the  port  x,  but  does  allow,  by  means  of  the  cavity  in  the  valve  v',  the  steam  previously 
admitted  through  this  port  x  to  escape  through  the  exhaust  port  x",  Figs.  1012  and 

1015,  to  the  open  air.     The  relief  thus  afforded  allows  the  live  steam  entering  the  port 
x'  to  force  the  piston  I  back  again  to  its  former  position,  as  shown  in  Fig.  1012,  which 
at  the  same  time  opens  the  intercepting  valve  d,  giving  free  communication  between 
the  high-  and  low-pressure  cylinders.     Since  the  backward  movement  of  the  intercept- 
ing valve  d  and  piston-rod  I,  Fig.  1012,  carries  back  also  the  lever  g,  which  allows  the 
valve  v  to  close,  thus  cutting  off  the  supply  of  live  steam  to  the  low-pressure  cylinder, 
the  engine  continues  to  operate  as  a  compound.    Owing  to  the  construction  of  the 
intercepting  valve  d,  Figs.  1011,  1012,  and  1020,  which  is  clearly  shown  in  the  illustra- 
tions, any  pressure  in  or  around  the  pistons  forming  the  valve  has  no  influence  what- 
ever upon  its  movements.     This  pressure  can  neither  start  the  valve  in  motion,  nor 
retard  its  movement  if  it  once  be  in  motion.     Therefore  its  movement  is  not  derived 
from  either  live  or  exhaust  steam  acting  upon  it.     The  only  use  made  of  the  exhaust 
steam  from  the  receiver  in  any  part  of  the  mechanism,  in  ordinary  operation,  is  to  move 
the  slide  valve  v',  Fig.  1012,  as  described.    In  the  event  of  the  intercepting  valve  being 
closed,  as  in  the  position  shown  in  Fig.  1021,  and  the  engine  running  with  the  throttle 
closed  as  in  the  case  of  a  locomotive  rolling  down  a  grade,  there  would  be  no  resisting 
pressure  to  cause  an  opening  of  the  intercepting  valve.     Unless  this  valve  should  be 
opened,  the  constantly  accumulating  gases  and  vapor  in  the  receiver  above  the  inter- 
cepting valve  would  finally  be  sufficient,  acting  as  back  pressure  on  the  H.  P.  piston, 
to  become  objectionable.     To  avoid  the  occurrence  of  such  a  condition,  the  auxiliary 
valve  v",  Fig.  1014,  is  employed.     This  receiver  pressure  is  carried  to  the  auxiliary 
valve  chamber  f,  Fig.  1014,  by  the  pipe  j/"  (Figs.  1011  and  1014),  and  acting  upon  the 
auxiliary  valve  v"  forces  it  to  the  opposite  end  of  its  chamber  from  that  shown  in  Fig. 
1014 ;  and  from  the  manner  in  which  it  seats  itself,  this  receiver  pressure  cannot  pass 
into  the  passage  p'.     The  above-mentioned  movement  of  the  valve  v"  uncovers  the  port 
«/',  Fig.  1014,  allowing  the  receiver  pressure  to  enter  the  valve  chamber  n  between  the  two 
pistons  t  and  t'.    At  the  same  time  this  receiver  pressure  enters  the  end  of  the  chamber 
n  through  the  port  x"'.     Since  the  pistons  t  and  t'  are  of  unequal  area,  the  same  press- 
ure between  the  two  pistons  and  on  the  opposite  side  of  the  larger  piston,  t',  would 
consequently  force  the  pistons  and  the  slide  valve  v1  to  assume  the  position  as  shown 
hi  Fig.  1012,  thus  uncovering  the  port  x'  allowing  the  receiver  pressure  to  enter  the 
cylinder  h,  and  by  its  action  force  the  piston  I  backward  as  shown  in  Fig.  1012,  thus 
opening  the  intercepting  valve  d. 

The  entire  arrangement  of  the  several  parts  of  the  mechanism  is  such  that  the 
intercepting  valve  is  moved  forward  and  backward  automatically,  and  at  proper  inter- 


MODERN  LOCOMOTIVE   CONSTRUCTION.  625 

vals,  to  meet  the  requiremeiits  of  any  or  all  circumstances  under  which  the  engine 
may  ho  operated. 

In  order  that  there  may  not  be  any  undue  slamming  of  the  intercepting  valve  as  it 
reaches  the  terminus  of  its  stroke  in  either  direction,  the  exhaust  cavity  of  the  valve  v', 
Figs.  1012  and  1015,  is  contracted  in  proper  proportion  to  the  volume  of  the  exhaust 
escaping  from  the  cylinder  /«,  so  as  to  form  compression  at  the  end  of  the  stroke,  thus 
allowing  the  intercepting  valve  to  travel  to  and  fro  at  a  reduced  speed  consistent  with 
its  weight  and  length  of  travel. 

The  above  is  all  the  extra  mechanism  that  is  required  for  a  compound  engine  of 
this  class.  In  fact,  to  change  a  simple  locomotive  to  a  compound,  all  that  is  usually 
done  is  to  take  off  one  cylinder,  replace  it  by  one  larger  in  diameter,  with  a  slide  valve 
and  steam-chest  necessarily  a  little  larger  than  used  in  the  simple  system,  and  attach 
the  intercepting  valve  with  its  auxiliaries. 

LOCOMOTIVES  WHICH  CAN  BE  WORKED  AS  COMPOUNDS  OE  NON-COMPOUNDS. 

641.  Figs.  1022,  1023,  1024  show  the  intercepting  valve  and  its  auxiliaries  for  a 
compound  locomotive,  which  can  be  changed  at  will  from  a  compound  to  a  non-com- 
pound, started  and  run  continuously  with  steam  directly  from  the  boiler  in  both  cylin- 
ders, each  doing  hah*  the  work,  and  each  exhausting  into  the  stack  and  atmosphere 
directly.  An  outside  view  of  this  engine  is  shown  in  Fig.  1025.  This  type  is  built  by 
the  Ehode  Island  Locomotive  Works,  Providence,  R.  I.  The  intercepting  valve  and 
auxiliaries  as  shown  in  the  following  illustrations  were  invented  by  Mr.  C.  H. 
Batchellor,  chief  draftsman  at  the  above  works. 

Fig.  1022  shows  a  front  section  of  the  intercepting  valve  at  ports  d  and  e,  also 
front  view  of  a  portion  of  the  receiver  with  exhaust  valve. 

Fig.  1023  shows  a  longitudinal  section  of  the  same  mechanism  while  running  com- 
pound, and  Fig.  1024  shows  the  same  while  running  simple. 

The  same  letters  in  the  different  views  represent  like  parts.  A,  represents  the 
intercepting  valve  casing ;  B,  the  reducing  valve ;  C',  the  oil  dash-pot ;  Z>,  the  pipe  from 
the  main  steam  pipe  to  the  intercepting  valve;  E,  the  receiver;  jP,  the  exhaust  valve; 
a  b  c,  the  intercepting  valve  pistons ;  d,  a  port  from  D  through  A ;  e,  a  port  from  A 
into  the  reducing  valve  B ;  f,  a  port  from  A  into  the  passage  to  L.  P.  steam-chest ;  HI, 
a  small  pipe  and  passage  connecting  hand  valve  in  cab  to  chamber  h ;  o,  port  through 
exhaust- valve  casing. 

The  operation  of  this  device  is  as  follows:  The  intercepting  valve  being  in  any 
position  and  the  exhaust  valve  F  closed  as  in  Fig.  1023,  then  on  opening  the  throttle 
steam  from  the  boiler  will  pass  into  the  H.  P.  cylinder  in  the  usual  manner,  and  also 
through  pipe  D  into  the  intercepting  valve  A,  causing  the  pistons  a  b  c  to  move  into 
the  position  shown  in  Fig.  1024.  In  this  position  the  receiver  is  closed  to  the  L.  P. 
cylinder  by  the  piston  c,  and  steam  from  1)  passes  through  the  ports  d  and  e  and 
reducing  valve  B  into  the  L.  P.  steam-chest,  the  pressure  being  reduced  from  the  boiler 
pressure  in  the  ratio  of  the  cylinder  areas.  The  pistons  a  I  c  are  so  proportioned  that 
they  will  automatically  change  to  the  compound  position  shown  in  Fig.  1023,  when  a 
predetermined  pressure  in  the  receiver  E  has  been  reached  by  the  exhaust  from  the 


MOI>l-:i!\  LOCOMOTIVE   CONSTRUCTION.  627 

H.  P.  cylinder.  The  engine  thus  starts  with  steam  in  both  cylinders,  and  changes 
automatically  to  a  compound  at  a  desired  receiver  pressure. 

The  engine  may  be  changed  from  a  compound  to  a  simple  system  at  any  time, 
at  the  will  of  the  engineer,  by  opening  the  valve  F  connecting  the  receiver  to  the 
exhaust  pipe,  allowing  the  exhaust  from  the  H.  P.  cylinder  to  escape  through  the 
nozzle  in  the  usual  manner. 

The  exhaust  valve  F  is  operated  as  follows :  The  small  pipe  m  leads  from  a  hand 
valve  in  the  cab  connecting  it  to  either  steam  or  atmosphere.  When  desiring  to  run 
compound,  m  is  in  connection  with  the  atmosphere,  the  receiver  steam  keeping  the 
valve  F  in  the  position,  as  shown  in  Fig.  1023.  To  run  simple  m  is  connected  to  the 
steam  which  will  hold  the  valve  F,  as  in  Fig.  1024,  the  ports  o  opening  E  to  the 
exhaust.  The  valve  F  takes  either  position  at  any  time  at  the  will  of  the  engineer. 

It  is  obvious  that  in  case  of  bad  conditions  of  starting,  the  engine  may  be 
operated  as  a  simple  one  by  opening  the  exhaust  valve  before  starting,  and  that  upon 
its  closure  the  pistons  a  b  c  will  automatically  take  the  compound  position. 

VALVE  GEAR  ADJUSTMENT  IN  TWO-CYLINDER  COMPOUNDS. 

642.  In  a  two-cylinder  compound  locomotive  the  division  of  work  is  equalized  by 
an  adjustment  of  the  valve  gear  so  as  to  cut  off  at  the  proper  time  in  each  cylinder. 
For  this  purpose  a  later  cut-off  is  frequently  given  in  the  low-pressure  cylinder  than 
in  the  high-pressure  cylinder.    Of  course,  for  a  correct  adjustment  of  the  valve  gear 
indicator-cards  will  be  required. 

The  difference  in  cut-offs  in  the  two  cylinders  is  frequently  obtained  by  making 
the  length  of  the  high-pressure  link-hanger  about  -^  of  the  lift  of  the  link  less  than 
that  of  the  low-pressure  link-hanger.  This  kind  of  adjustment  is  of  course  suitable 
only  for  engines  which  run  mostly  in  forward  gear ;  for  others,  which  run  as  much  in 
one  direction  as  in  the  other,  this  adjustment  is  not  suitable.  In  the  latter  cases  the 
required  cut-offs  can  be  obtained  by  reducing  the  angular  advance  of  the  low-pressure 
eccentric,  and  also  reducing  to  a  corresponding  degree  the  outside  lap  of  the  valve  for 
the  same  cylinder. 

STEAM-PORT  AREA  FOE  COMPOUND  LOCOMOTIVE. 

643.  The  steam-port  area  for  the  high-pressure  cylinder  is  found  by  the  same  rule 
as  given  in  Art.  43.     This  rule  can  bo  given  in  another  form,  so  that  the  use  of  table 
given  in  Art.  43  is  not  necessaiy ;  thus : 

RULE  129. — Multiply  the  speed  of  piston  in  feet  per  minute  by  the  square  of  the 
diameter  in  inches,  and  divide  the  result  by  7639 ;  the  quotient  will  be  the  area  in 
square  inches  of  the  steam  port  in  the  H.  P.  cylinder.  Expressing  this  rule  by 
symbols,  we  have 

(Diameter  of  cylinder  in  inches)2  x  speed  of  piston  in  feet 

-  =  area  of  steam  ports. 
7639 

The  ratio  of  width  to  length  of  steam  port  is  the  same  as  given  in  Art.  43. 

For  finding  the  steam-port  area  for  the  low-pressure  cylinder,  we  IKIV 

RULE  130. — Multiply  the  piston  speed  in  feet  per  minute  by  the  square  of  the 


628  MODERN  LOCOMOTIVE   CONSTRUCTION. 

diameter  of  the  cylinder  in  inches,  and  divide  the  product  by  8800  ;  the  quotient  will 
be  the  stearn-port  area  for  the  low-pressure  cylinder.  Expressing  this  rule  by  symbols, 
we  have 

(Diameter  of  cylinder  in  inches)2  x  speed  of  piston  in  feet 

~~8800~~  =  area       steam  port. 

From  this  it  will  be  seen  that  the  steam-port  area  in  the  L.  P.  cylinder  is  not  as 
great  in  proportion  to  the  piston  area  as  in  the  H.  P.  cylinder. 

EXAMPLE  168.  —  What  should  be  the  dimensions  of  the  steam  ports  in  a  compound 
locomotive  with  cylinders  17  and  24  inches  diameter,  piston  speed  700  feet  per  minute? 

For  the  area  of  steam  ports  in  the  H.  P.  cylinder,  we  have,  according  to  rule, 

17  2  x  700 

=  26.48  square  inches. 


If  we  now  decide  to  make  the  width  of  the  steam  port  If  inches,  then  the  length 

will  be 

26.48 

..  £05  =  16.29,  say  16^  inches. 

For  the  area  of  the  steam  ports  in  the  L.  P.  cylinder,  we  have,  according  to  rule, 

242  x  700 

=  45.81  square  inches. 


If  we  now  decide  to  make  the  length  of  the  steam  port  22  inches,  then  we  have  for 
its  width 

-~-  =  2.08,  say  2  inches. 

DIAMETER  OF  CYLINDERS. 

644.  There  exists  a  difference  in  opinion  in  regard  to  the  proper  diameters  of 
cylinders  in  a  two-cylinder  compound,  and  consequently  several  rules  for  obtaining 
them  have  been  recommended.  We  believe  that  the  simplest  way,  and  which  will 
probably  give  as  satisfactory  results  as  any,  is  to  find  the  diameter  of  the  low-pressure 
cylinder  by  Eule  115,  using  25  per  cent,  of  the  boiler  pressure  for  the  mean  effective 
pressure  p;  then  make  the  area  of  the  H.  P.  cylinder  one  half  that  of  the  L.  P. 
cylinder.  According  to  this  the  volume  of  the  L.  P.  cylinder  is  twice  that  of  the  H.  P. 
cylinder,  provided  the  stroke  is  the  same  in  both  cylinders.  This  ratio  is  not  always 
adopted;  frequently  it  is  somewhat  greater,  making  the  volumes  as  1  to  2.2,  and 
occasionally  we  find  this  ratio  somewhat  less  than  1  to  2. 

EXAMPLE  169. — What  should  be  the  diameters  of  cylinders  for  a  compound  loco- 
motive having  a  tractive  force  of  11,472  pounds ;  stroke  of  piston,  24  inches ;  diameter 
of  driving  wheels,  61  inches;  boiler  pressure,  170  pounds?  For  the  mean  effective 
pressure  we  take  170  x  .25  =  42.5  pounds.  The  formula  given  in  Art.  594  for  finding 
the  diameter  of  the  cylinder  is 

riTD 

p  x  s' 
hence  the  diameter  of  the  low-pressure  cylinder  will  be 

11472  x  61 

24    =  26  inches  (fraction  omitted). 


MODERN  LOCOMOTIVE  CONSTRUCTION.  629 

The  area  of  a  26-inch  piston  is  530.93  square  inches.  If  we  now  decide  to  make  the 

530.93 
area  of  H.  P.  cylinder  one  half  of  that  of  L.  P.  cylinder,  then  we  have  — „ —  =  265.46 

square  inches  for  the  area  of  H.  P.  cylinder,  and  the  corresponding  diameter  will  be 
18 1  inches. 

Referring  to  Table  5,  page  17,  we  find  that  a  non-compound  eight-wheeled 
passenger  engine  with  the  same  tractive  force,  same  size  of  wheels,  but  with  a  mean 
effective  pressure  of  90  pounds  per  square  inch,  has  cylinders  18  x  24  inches. 

If  in  the  foregoing  example  the  weight  on  drivers  had  been  given  instead  of  the 
tractive  force,  we  should  have  divided  the  weight  on  drivers  by  5,  which  would  have 
given  us  the  tractive  force,  and  then  proceeded  as  before. 

The  foregoing  method  of  finding  the  diameters  of  cylinders  for  a  compound  engine 
applies  to  those  only  in  which  the  area  of  the  L.  P.  cylinder  is  about  twice  that  of  the 
H.  P.  cylinder. 

VOLUME  OF  RECEIVER. 

The  volume  of  the  receiver  should  not  be  less  than  that  of  the  H.  P.  cylinder ;  in 
fact  a  receiver  volume  somewhat  greater  than  this  will  give  better  results,  as  such  will 
reduce  the  fluctuations  of  back  pressure  in  the  H.  P.  cylinder. 

FOUR-CYLINDER  COMPOUND  LOCOMOTIVES. 

645.  The  four-cylinder  compounds  used  to  a  great  extent  in  the  country  are  built 
by  the  Baldwin  Locomotive  Works  of  Philadelphia,  and  patented  by  Mr.  Samuel  M. 
Vauclain,  superintendent  of  these  works. 

The  general  construction  of  boiler,  frames,  valve  gear,  driving  and  truck  wheels, 
is  the  same  as  in  an  ordinary  non-compound ;  the  difference  between  these  engines  is 
entirely  in  the  cylinders,  pistons,  cylinder  cocks,  crosshead,  and  valves,  as  shown  in 
Figs.  1026,  1027 ;  the  latter  shows  the  construction  of  the  main  valve,  and  Fig.  1028 
shows  the  starting  valve. 

A  piston  valve,  shown  in  section  in  Fig.  1027,  is  used  for  distributing  the  steam  in 
the  high-  and  low-pressure  cylinders  which  are  placed  on  each  side  of  the  engine ;  the 
flow  of  steam  is  clearly  indicated  by  the  arrows.  The  steam  first  enters  the  H.  P. 
cylinder  and  propels  the  piston  in  it ;  on  the  return  stroke  it  passes  through  the  central 
chamber  in  the  valve  to  the  opposite  end  of  the  L.  P.  cylinder,  where  it  expands  and 
propels  the  large  piston,  and  in  the  next  stroke  passes  through  a  circular  groove  in 
the  center  of  the  valve,  and  is  discharged  through  the  exhaust  port  and  exhaust  pipe 
in  the  usual  manner.  The  same  operation  takes  place  in  both  ends  of  the  cylinders. 

The  centers  of  the  high-  and  low-pressure  cylinders  are  placed  in  one  vertir.-il 
plane,  and  the  steam-chest  is  placed  on  the  inside  of  the  cylinders,  as  shown  in  Fig. 
1026.  Both  piston-rods  are  connected  to  one  crosshead,  and  the  valve-rod  connects  to 
the  upper  rocker-arm  in  the  usual  manner. 

The  piston  valve  being  balanced,  it  takes  comparatively  very  little  force  to  reverse 
the  engine;  and  the  pressure  of  the  valve  against  its  seat  being  reduced,  it  is  not  so 
liable  to  run  dry  as  an  ordinary  slide  valve  working  under  the  same  high  steam 


630 


MODERN   LOCOMOTirE    CONSTRUCTION. 


COMBINATION  VALVE  LEVER 

LOCATED  IN  CAB 

FOR  OPERATING  CVL  COCK  &  STARTING 
VALVC. 


pressure,  and  therefore  the  piston  valve  is  well  adapted  for  compound  engines  in  which 
high  steam  pressure  is  used. 

(J4G.  Fig.  1028  shows  a  section  of  the  starting  valve.  It  is  attached  to  the  bottom 
of  the  L.  P.  cylinder,  as  shown  in  Fig.  1029,  and  is  operated  in  the  same  way  as 
ordinary  cylinder  cocks.  It  consists  of  a  casing  with  a  piston  valve  inside.  The 
no/./les  A  and  />  are  in  communication  with  the  L.  P.  cylinder,  and  the  nozzles  C,  D 
are  connected  by  pipes  to  H.  P.  cylinder.  The  piston  valves  //and  L  fully  cover  the 
openings  to  the  L.  P.  cylinder,  while  the  piston  K  covers  the  opening  from  one  side  of 
the  H.  P.  cylinder;  steam  is  therefore  shut  off  from  communication  between  the  two 
cylinders  and  from  the  condensation  vents  E  and  G. 

In  this  position  of  the  piston  valves,  the  lever  in  the  cab  which  operates  them 
occupies  the  position  as  indicated  at  P,  Fig.  1030.  When  this  lever  occupies  the 
position  of  /'2  the  valve 
is  moved  half  stroke, 
still,  however,  covering 
the  L.  P.  exits,  but  un- 
covering the  port  />, 
throwing  the  two  ports 
C  and  D  in  communi- 
cation, permitting  steam 
from  the  inlet  end  of  the 
H.  P.  cylinder  to  travel 
through  the  starting 
valve  to  the  exhaust  end 
of  the  H.  P.  cylinder, 
and  consequently  into 
the  inlet  of  the  L.  P. 
cylinder,  thus  mixing 
live  steam  with  the  par- 
tially expanded  steam. 
When  the  lever  in  the 
cab  occupies  the  position  of  P3  the  piston  valve  has  traveled  to  its  extreme  position 
and  opened  the  vents  E,  G  of  the  low-pressure  cylinder,  while  leaving  the  live  steam 
inlet  also  open.  The  starting  valve  may  be  opened  without  opening  the  vents  which 
act  as  cylinder  cocks,  but  the  latter  cannot  be  opened  without  admitting  live  steam 
into  the  L.  P.  cylinder.  Further,  the  rigging  in  the  cab  is  so  arranged  that  the 
starting  valve  may  be  kept  open  at  will  for  any  length  of  time,  with  the  reverse  lever 
in  any  position.  This  makes  an  extremely  simple  starting  valve  and  enables  the 
engineer  to  admit  live  steam  into  the  L.  P.  cylinder  at  will. 

SI/E  OF   OYLINM.KS    IN    FOUR-CYLINDER  LOCOMOTIVES. 

647.  The  formula  used  by  the  Baldwin  Locomotive  Works  for  finding  the  size  of 
cylinders  is  as  follows: 

Diameter  of  L.  P.  cylinder  =  J~lz*«}«  ""  ''rive^ta  pouiidg  x  diameter  of  driv.-n.  in  in,-h,?_t 

1     luiili-r  ]ii-i-isnri.  in  |ii>iind<  |»-r  s'|ii:in-  inch   x   xtroke  in  inches  x  2.7 


632  MODERN  LOCOMOTIVE  CONSTRUCTION. 

In  ordinary  language  this  reads : 

KULE  131. — Multiply  the  weight  on  drivers  in  pounds  by  the  diameter  of  the 
drivers  in  inches;  call  this  product  A.  Multiply  the  boiler  pressure  in  pounds  per 
square  inch  by  the  stroke  in  inches  and  by  the  constant  2.7 ;  call  this  product  B. 
Divide  product  A  by  product  B,  and  extract  the  square  root  of  the  quotient.  The 
result  will  be  the  diameter  of  the  L.  P.  cylinder. 

The  area  of  the  H.  P.  cylinder  is  made  as  nearly  as  possible  one  third  of  the  area 
of  the  L.  P.  cylinder. 

EXAMPLE  170. — What  should  be  the  diameters  of  the  cylinders  of  a  four-cylinder 
compound,  the  weight  on  drivers  is  57,360  pounds;  stroke  of  pistons,  24  inches; 
diameter  of  drivers,  61  inches ;  steam  pressure,  180  pounds  ! 

Diameter  of  L.  P.  cylinder  =  Vi80~>   ''>4  x  27  =  17'3'  say  17^  incnes- 
The  area  of  this  cylinder  is  233.7055  square  inches,  consequently  the  area  of  the 

233  7055 

H.  P.  cylinder  should  be  — '-~  -  =  77.9018  square  inches ;  the  corresponding  dia- 
meter is  10  inches  very  nearly.  Hence  the  high-pressure  cylinder  should  be  10  inches 
diameter  and  the  low-pressure  cylinder  17 J  inches  diameter. 

Up  to  this  time  about  200  engines  of  this  class  have  been  built.  Besides  these  a 
great  number  of  two-cylinder  engines  have  also  been  built  by  various  locomotive 
builders,  and  the  average  reports  in  regard  to  results  as  far  as  we  can  learn  are  very 
favorable,  while  a  few  are  very  contradictory.  But  the  indications  are  that  as  com- 
pound locomotives  are  more  generally  used,  and  improvements  made  as  experience 
may  suggest,  they  will  occupy  a  prominent  place  in  railroad  service. 


THE   END. 


INDEX. 


Absolute  initial  steam  pressure,  597. 

Absolute  steam  pressure,  .V.I7. 

Action  of  eccentric,  80. 

Action  of  exhaust  steam,  502. 

Action  of  spring  equalizing  levers,  400. 

Adhesion,  7. 

Adjustment  of  valve  gear  in  two-cylinder  compounds,  627. 

Admission  line,  607. 

Admission  of  steam,  to  find  the  point  of,  60. 

Adoption  of  eccentric,  reason  for  the,  38,  80. 

Advance,  linear,  of  valve,  45. 

Advantage  gained  by  dividing  the  steam  passage  into 

two  branches,  22. 

Advantage  of  cutting  off  early,  609. 

Advantage  of  the  present  management  of  locomotives,  2. 
Advantages  claimed  for  Allen  valve,  69. 
Advantages  of  placing  valve  in  central  position,  44. 
Advantages  of  single  and  double  exhaust  pipes,  506. 
Air  admitted  through  hollow  brick  arch,  temperature  of, 

477. 

Air  chambers  for  pumps,  390. 
Air  chambers  for  pumps,  capacity  of,  390. 
Air  chambers  for  pumps,  position  of,  394. 
Allen  valve,  68. 

Allowance  for  shrinkage  in  driving-wheel  tires,  217. 
Allowance  for  wear  and  re-planing  of  guides,  167. 
American  locomotives,  shifting  link  for,  88. 
Amount  of  axle  and  rolling  friction,  5. 
Amount  of  counterbalance,  228. 
Amount  of  piston  clearance,  testing  the,  125. 
Amount  of  throw  of  eccentric's.  103. 
Amount  of  weight  that  rails  can  bear,  8. 
Angle  of  safety-valve  bearing  surface,  371. 
Angle  of  surface  in  throttle  valves,  343. 
Angular  advance  affected  by  lap,  45. 
Angular  advance  for  shifting  links,  91. 
Angular  advance  for  stationary  links,  to  find,  91. 
Angular  advance  of  eccentric.  4f>. 

Angular  advance  of  eccentric  for  stationary  link,  88,  91. 
Angular   advance   of   eccentrics;    rocker-arms  of   equal 

lengths,  114. 
Angular  advance  of  eccentrics ;  rocker-arms  of  unequal 

lengths,  114. 
Angular  advance  of  eccentrics;  rocker  employed,  center 

line  of  motion  of  valve  gear  not  coinciding  with  that 

of  piston,  116. 

Angular  advance  of  eccentric,  to  find,  98.  99. 
Angular  advance  of  eccentric  when  rocker  is  used,  111. 
Angularity  of  the  eccentric-rods,  complicating  influence 

of,  36". 

Annealing,  strength  of  plates  restored  by,  451. 
Anthracite  coal  burning  engines,  424. 
Apparent  strength  of  plates.  4.11. 
Appearance  of  compounds.  (J'Jti. 
Appearance  of  locomotives,  change  in,  3. 
Approximate'  rule  for  finding  thrust  against  guides,  161. 
Arc.  center  of  gravity  of  an.  L'.~>7. 
Arc,  definition  of  radius  of  link,  89. 


Arc  described  by  the  end  of  reverse  lever,  length  of,  109. 

Arc  for  eccentric-rod  pins,  100. 

Arc,  length  of  radius  for  link,  90. 

Arc  of  link,  definition  of,  99. 

Arc  on  throttle  valve-rod  jaw,  length  of,  363. 

Arcs  for  reverse  lever,  109. 

Arcs  for  reverse  lever,  notches  in,  109. 

Arcs  for  reverse  lever,  to  lay  off  notches  on,  133. 

Area,  cross-sectional,  of  crown-bar  braces,  466. 

Area,  crosfc-sectional,  of  engine-frame  braces,  191. 

Area,  cross-sectional,  of  exhaust  pipes,  507. 

Area,  cross-sectional,  of  rectangular  part  of  throttle 
pipe,  349. 

Area,  cross-sectional,  of  slab-frame  brace,  197. 

Area  of  central  side-rod  for  consolidation  engines,  cross- 
sectional,  312. 

Area  of  crank-pins,  projected,  318. 

Area  of  crosshead-pin,  projected,  174. 

Area  of  equalizing  lever  fulcrum,  cross-sectional,  409. 

Area  of  front  and  rear  side-rod  pins  for  Mogul  and  ten- 
wheeled  engines,  projected,  .'i.'W. 

Area  of  grate  surface,  hard  coal,  426. 

Area  of  grate  surface,  soft  coal,  422. 

Area  of  irregular  surfaces,  254. 

Area  of  knuckle-joint  pin  in  side-rods,  projected,  340. 

Area  of  orifices  in  exhaust  nozzles,  507. 

Area  of  pump  plungers,  full  stroke,  cross-sectional,  396. 

Area  of  rim  of  wheel,  cross-sectional,  256. 

Area  of  safety  valve,  376,  379. 

Area  of  side-rod  pins  for  consolidation  engines,  pro- 
jected, 336. 

Area  of  side-rods,  cross-sectional,  308. 

Area  of  sliding  surface  of  crosshead  gibs,  163. 

Area  of  smallest  transverse  section  of  main-rod,  301. 

Area  of  spring  hangers,  cross-sectional,  410. 

Area  of  stacks,  cross-sectional,  502. 

Area  of  steam  and  exhaust  ports,  26. 

Area  of  steam  passages,  31. 

Area  of  steam  pipes,  30. 

Area  of  steam  pipes,  to  compute,  30. 

Area  of  steam  pipes,  to  compute  with  aid  of  table,  31. 

Area  of  steam  ports  in  compounds,  627. 

Area  of  steam  port,  to  compute,  27. 

Area  of  surface  of  counterbalance,  249. 

Area  of  tubes,  cross-sectional,  438. 

Areas,  table  of  proportional  steam  pipe,  31. 

Areas,  table  of  proportional  steam  port,  28. 

Ann  of  a  force,  or  lever  arm,  225. 

Arm  of  counterbalance,  242. 

Arms  of  rockers,  length  of,  76. 

Arms  of  rockers,  thickness  of,  77. 

Anns  of  rockers,  width  of,  77. 

Arms  on  lifting  shaft,  dimensions  of.  107. 

Arms  on  lifting  shaft,  location  of,  107. 

Arrangement  of  simple  valve  gear.  :;ii. 

Arrangement  of  steam  ways  ami  ports  in  cylinder  and 
saddle.  HII. 

Arrangement  of  tubes.  442. 

Asbestos  paper  for  cylinder  lagging,  24. 

Ash-pan,  cast-iron,  494. 


633 


634 


INDEX. 


Ash-pan  dampers,  493,  494. 

Ash-pans  for  consolidation  engines,  494. 

Ash-pans  for  eight-wheeled,  Mogul,  and  ten-wheeled  en- 
gines, 491. 

Ash-pans  with  blowing-out  arrangement,  493. 

Ash-pang  with  side  doors,  493. 

Atmospheric  line,  602,  607. 

Attachments  for  safety  valves,  371. 

Axle  and  rolling  friction,  5. 

Axle  and  rolling  friction,  methods  of  finding,  6. 

Axle-box  brasses,  204. 

Axle-box  brasses,  Babbitt  metal  in,  205. 

Axle-box  brasses  for  tender  trucks,  width  of,  581. 

Axle-box  brasses,  octagonal,  objections  to,  204. 

Axle-box  brasses,  preparation  of,  for  lead  lining,  577. 

Axle-box  brasses,  pressure  for  forcing  into  box,  205. 

Axle-boxes,  blocking  up  the,  126. 

Axle-boxes,  classification  of,  203. 

Axle-boxes,  designing  of  driving,  206. 

Axle-boxes  for  engine  trucks,  548. 

Axle-boxes  for  main  axles,  203. 

Axle-boxes  for  Mogul  and  consolidation  engines,  212. 

Axle-boxes  for  tender  trucks,  575. 

Axle-boxes  for  tender  trucks,  brasses  for,  575. 

Axle-boxes  for  tender  trucks,  lead  lining  for  brasses,  577. 

Axle-boxes  for  tender  trucks,  wedges  for,  575. 

Axle-boxes,  play  between  wheels  and,  203. 

Axle-boxes,  position  of  center  lines  on  frames,  190. 

Axle-boxes,  proportions  of,  205. 

Axle-boxes,  vertical  clearance  in  engine  pedestals,  189. 

Axle-boxes,  vertical  movement  of,  189. 

Axle-boxes,  width  of  driving,  208. 

Axle-box  for  tender  trucks,  cotton  or  woolen  waste  in, 
578. 

Axle-box  for  tender  trucks,  cover  for,  578. 

Axle-box  for  tender  trucks,  design  of,  578. 

Axle-box  for  tender  trucks,  dust  guard  for,  578. 

Axle-box  for  tender  trucks,  standard,  579. 

Axle-box  for  tender  trucks,  types  of,  579. 

Axle-box  for  tender  trucks  with  pedestals,  579. 

Axle-box  for  tender  trucks  without  collars  on  axle,  579. 

Axle-box  for  tender  trucks  without  wedge,  578. 

Axle-box  oil  cellars,  204. 

Axle-box,  Pennsylvania  E.  R.,  main,  413. 

Axle-box,  pockets  in,  204. 

Axle,  distance  from  fire-box  to  center  of,  82. 

Axle  journal,  diameter  of  driving,  209. 

Axle  journals  for  eight-wheeled,  Mogul,  ten-wheeled,  and 
consolidation  engines,  210. 

Axle  journals  for  engine  trucks,  549. 

Axle  journals  for  tender  trucks,  580. 

Axle  journals  for  tender  trucks,  diameter  of,  581. 

Axle  journals  for  tender  trucks,  pressure  on,  581. 

Axle  journals  for  tender  trucks,  to  compute  dimensions 
of,  582. 

Axle  journals  for  tender  trucks,  to  compute  load  they 
can  bear,  582. 

Axle,  key  ways  in,  for  eccentric,  81. 

Axle,  position  of,  for  laying  out  the  valve  gear,  119. 

Axles,  distance  between  centers  of  engine  truck,  538. 

Axles,  distance  between,  in  consolidation  engines,  188. 

Axles,  distance  between,  in  eight-wheeled  engines,  188. 

Axles,  distance  between,  in  Mogul  engines,  188. 

Axles,  distance  between,  in  ten-wheeled  engines,  188. 

Axles,  driving,  214. 

Axles  for  tender  and  engine  trucks,  difference  between 
pressure  on,  582. 

Axles  for  tender  trucks,  standard,  581. 

Axles,  greatest  distance  between  centers  of,  188. 

Axles  in  soft-coal  or  wood-burning  engines,  space  re- 
quired between,  188. 

Axles,  under  hard-coal  burners,  position  of,  189. 

Axle,  to  find  position  of  eccentrics  on,  132. 


Babbitted  wings  on  crossheads,  151. 
Babbitt  metal  for  rod  brasses,  315. 


Babbitt  metal  in  driving  axle-box  brasses,  205. 

Back  ends  of  engine  frames  suitable  for  foot-plates,  187. 

Backpressure,  602. 

Back  pressure  line,  or  line  of  counter-pressure,  607. 

Backs  of  flanges  of  tires,  distance  between,  266. 

Backward  eccentric,  89. 

Backward  eccentric-rod,  89. 

Backward  stroke,  89. 

Balanced  slide  valves,  66. 

Balanced  slide  valves,  hole  in  top  of,  67. 

Balanced  slide  valves,  Richardson's,  67. 

Balanced  slide  valves,  springs  for,  66. 

Balance  for  safety-valve,  spring,  362,  369. 

Baldwin  Locomotive  Works,  consolidation  engine  built 

by  the,  589. 
Baldwin  Locomotive  Works,  ten-wheeled  engine  built  by 

the,  589. 

Ball  joint  between  dry  pipe  and  front  flue  sheet,  347. 
Ball  joint  between  dry  pipe  and  T-pipe,  350. 
Ball  joint  between  throttle  and  dry  pipes,  347. 
Ball  joint  for  steam  pipes,  33. 

Batchellor,  C.  H.,  intercepting  valves  for  compounds,  625. 
Bearers  for  water  grate,  489. 
Bearing  surface  of  safety  valves,  angle  of,  371. 
Bell  crank  for  throttle  valve,  343. 
Bells,  construction  of,  514. 
Bells,  stand  and  yoke,  513. 
Belly  brace,  523. 

Belpaire  boiler,  form  of  crown  sheet  in,  463. 
Belpaire  boiler,  stay  bolts  in,  464. 
Blades  in  driving  wheel  springs,  number  of,  415. 
Blades  in  driving-wheel  springs,  thickness  of,  415. 
Block  link,  clearance  between  ends  of  link  and,  130. 
Blocks,  copper  strips  between  guides  and,  152. 
Blocks  for  guides,  149. 
Blower  valves,  558. 

Bodies,  common  center  of  gravity  of  any  two,  240. 
Boiler  braces,  468. 

Boiler,  capacity  of  smoke-boxes  on,  473. 
Boiler,  distance  between  centers  of  stay  bolts,  480. 
Boiler  domes,  470. 
Boiler,  extension  fronts,  472. 
Boiler  fastenings  to  frames,  482. 
Boiler,  gusset  plates  in,  470. 
Boiler  lagging,  extension  of,  374. 
Boiler  lagging,  thickness  of,  366. 
Boiler,  manner  of  staving  back  head  of,  469. 
Boiler,  manner  of  staying  front  tube  sheet,  469. 
Boiler  pads,  482.     . 

Boiler  pads,  clearance  between  equalizing  lever  and,  402. 
Boiler  .plates,  formulas  for  computing  the  thickness  of, 

462. 

Boiler  plates,  fun-owing  in,  459. 
Boiler  pressure,  formulas  for  computing  the,  462. 
Boiler,  reinforcing  plates  for  pads  on,  483. 
Boiler  rivets,  stress  in,  452. 
Boilers,  417. 

Boilers,  anthracite  coal  burning,  421,  424. 
Boilers,  area  of  grate  surface  for  hard  coal,  426. 
Boilers,  area  of  grate  surface  for  soft  coal,  422. 
Boilers,  arrangement  of  tubes,  442. 
Boilers,  beading  of  tubes,  437. 
Boilers,  Belpaire  fire-box,  432. 
Boilers,  bituminous  coal  burning,  421. 
Boilers,  blow-off  cocks  in,  418. 
Boilers,  butt  joints  in,  460. 
Boilers,  classification  of,  421. 
Boilers,  copper  ferrules  on  tubes,  438. 
Boilers,  cross-sectional  area  of  tubes,  438. 
Boilers,  crown-bar  bolts,  465. 
Boilers,  crown  bars,  464. 
Boilers,  crown  bars,  strength  of,  467. 
Boilers,  dents  in  side  of  fire-box,  437. 
Boilers,  depth  of  fires,  424. 
Boilers,  depth  of  hard  and  soft  coal  burning  fire-boxes, 

428. 

Boilers,  diameter  and  design  of,  443. 
Boilei-s,  diameter  of  rivets,  450. 


IXDEX. 


635 


Boilers,  diameter  of  tube  affects  heating  surface,  441. 

Boilers,  distance  of  rivets  from  edge  of  plate,  448. 

Boilers,  lire-box  by  Mr.  J.  HeiicUleii,  424. 

[{oilers,  tire-lxix  ring,  434. 

Boilers,  fire-door  opening,  form  of,  436. 

Boilers,  Hue  sheets,  thickness  of,  4:!4. 

Boilers,  form  of  crown  sheets  in,  4.'!^,  463. 

Boilers,  furnace  sheets,  thickness  of,  434. 

Boilers,  grate  surface.  421. 

Boilers,  greatest  width  of  fire-box,  427. 

Boilers,  hand  holes  in,  417. 

Boilers,  heating  surface,  445. 

Boiler  shell,  thieknessof,  460. 

Boiler,  short  smoke-boxes  on,  470. 

Boilers,  inclination  of  furnace  door  and  side  sheets,  432. 

Boilers,  large  diameter  of  tube  will  stiffen  them,  441. 

Boilers,  length  of  tuoes.  441'. 

Boilers,  limit  of  si/.e  of  fire-box,  422. 

Boilers,  mud  drums  on,  417. 

Boilers,  number  of  tubes,  440. 

Boilers,  position  of  blow-off  cocks  in,  418. 

Boilers,  position  of  tire-box  relative  to  axles,  421. 

Boilers,  position  of  hand  holes  in.  4'_'1. 

Boilers,  radial  stay  bolts  for  crown  sheet,  432,  463. 

Boilers,  rate  of  evaporation  on  inclined  sheets,  432. 

Boilers,  ratio  of  diameter  to  length  of  tubes,  441. 

Boilers,  reinforcing  plates  for  mud  plugs  in,  417. 

Boilers,  riveted  joints,  446. 

Boilers,  rivets  in  furnace  sheets,  435. 

Boilers,  single-riveted  lap  joints,  447. 

Boilers,  sloping  crown  sheet,  430. 

Boilers,  space  between  tubes,  443. 

Boiler  stay  bolts,  463. 

Boiler  stay  bolts,  hollow,  463. 

Boiler,  steam  chamber  for  receiving  valves  on,  561. 

Boilers,  steam  space,  443. 

Boilers,  stress,  longitudinal,  447. 

Boilers,  stress,  transverse,  447. 

Boilers,  tie  rod  in,  469. 

Boilers,  transverse  braces,  467. 

Boiler,  stress  in  oblique  braces,  481. 

Boiler,  stress  in  stay  bolts,  479. 

Boilers,  tubes,  4ii7. 

Boiler  supports,  522. 

Boilers,  width  of  water  space,  434. 

Boilers,  wood-burning  fire-box,  430. 

Boiler,  Wootten,  478. 

Bolts,  allowable  stress  per  square  inch  of  section  of  rod, 

275. 

Bolts  and  wedges  for  engine  frame,  proportion  of,  186. 
Bolts,  factor  of  safety  for  rod,  275. 
Bolts  for  engine  frames,  201. 
Bolts  for  guides,  168. 

Bolts  for  main-  and  side-rods,  taper  of,  267,  299. 
Bolts  for  segments  of  counterbalance,  242. 
Bolts  for  side-rods,  eight-wheeled  engines,  cross-sectional 

area  of,  297. 

Bolts  for  side-rods,  eight-wheeled  engines,  stress  in,  297. 
Bolts  for  side-rods  for  consolidation  engines,  diameter 

of,  298. 

Bolts  for  side-rods  for  Mogul  engines,  diameter  of,  298. 
Bolts  for  side-rods  for  ten-wheeled  engines,  diameter  of, 

298. 

Bolts  for  side-rods,  limited  number  of,  296. 
Bolts  for  side-rods,  number  and  diameter  of,  267. 
Bolts  for  side-roils,  strength  of,  296. 
Bolts  in  cylinder  heads.  2X 

Bolts  in  engine-frame  pedestal  caps,  number  of,  194. 
Bolts  in  main-rod,  number  and  diameter  of.  Ji',7. 
Bolts  in  side-rods  and  main-rods,  position  of.  I'il'.i. 
Bolts  in  skeleton  links,  difficulty  with.  lo:!. 
Bolts,  longitudinal,  through  tender  frames,  563. 
Bolts,  objection  to.  iii  cylinder  heads,  24. 
Bolts,  shearing  stress  in  rod,  27± 
Bolts,  stress  in  cylinder  head.  24. 
Bolts  through  cylinder  saddle  and  front  splice.  199. 
Bolts  through  frame  and  splice,  diameter  of,  198. 
Bolts  through  frame  and  splice,  stress  in,   MIS. 


Bolts  through  front  straps  of  main-rod,  diameter  of,  279. 

Bolts  through  rear  strap  of  main-rod,  diameter  of,  276. 

Bolts  through  rear  strap  of  main-rod,  number  of,  276. 

Bolts,  to  find  the  diameter  of  cylinder  head,  23. 

Boxes  for  sand,  510. 

Box  link,  99. 

Boyle's  law,  598. 

Braces  for  crown  bars,  466. 

Braces  for  engine  frames,  equal  depth  throughout,  196. 

Braces  for  engine  frames,  stress  per  square  inch,  191. 

Braces  for  engine  truck,  535. 

Braces  for  foot-plate,  524. 

Braces,  frame,  in  a  consolidation  engine,  524. 

Braces,  frame,  in  a  Mogul  engine,  524. 

Braces,  frame,  in  an  eight-wheeled  engine,  523. 

Braces  from  frames  to  boilers,  522. 

Bracket  lamp,  for  steam  gauge,  302. 

Brackets  for  running  boards,  531. 

Brake,  manner  of  attaching  Wcstinghouse,  585. 

Brakes,  gear  for  tender  trucks,  504,  571. 

Brass  casing  on  throttle  valve-rod,  355. 

Brass  driving  axle-boxes,  204. 

Brasses,  flanges  on  rod,  316. 

Brasses  for  driving  axle-boxes,  204. 

Brasses  for  driving  axle-boxes,  octagonal,  204. 

Brasses  for  driving-boxes,  heating  of,  206. 

Brasses  for  main  crank-pin,  caps  on,  316. 

Brasses  for  rods,  315. 

Brasses  for  solid-ended  side-rods,  326. 

Brasses  for  tender-truck  axle-boxes,  575. 

Brasses  for  tender  trucks,  lead  lining  for,  577. 

Brasses  in  valve-rod  end,  79. 

Brasses,  metal  in  rod,  317. 

Brasses,  pressure  for  forcing  them  into  box,  205. 

Brasses,  thickness  of  metal  in  rod,  317. 

Brass  gibs  for  crossheads,  151. 

Brass  packing  rings  for  pistons,  143. 

Brass  packing  rings  for  pistons,  width  and  depth  of,  144. 

Brass  ring  inside  of  stuffing-box,  179. 

Breadth  and  length  of  crosshead  gibs,  164. 

Breadth  of  exhaust  ports,  29. 

Breadth  of  guides,  166. 

Breadth  of  link,  105. 

Breadth  of  steam  and  exhaust  ports,  26. 

Breadth  of  steam  ports,  29. 

Brick  arch,  473,  477. 

Bridges  in  cylinder,  22. 

Bridges  in  cylinder,  thickness  of,  25. 

Bridges,  weight  on  driving  wheels  limited  by,  8. 

Buckling  of  side-rods,  .'!(><>. 

Built-up  crosshead  for  four  slides,  152. 

Built-up  engine  frame,  195. 

Built-up  pistons,  138. 

Bumper  beams,  518. 

Bumper  beams,  computation  of  breadth  of,  520. 

Bumper  beams,  stress  in,  520. 

Bumper  brace  for  extension  fronts,  522. 

Bumper  brace  for  short  smoke-boxes,  522. 

Bumper  plate  or  sheet,  522. 

Bushing  for  eccentric-rod  pins,  103. 

Bushing  for  eccentric-rod  pins,  loose,  103. 

Bushing  for  eccentric-rod  pins,  pressure  for,  103. 

Bushing  for  eccentric-rod  pins,  thickness  of,  103. 

Bushing  for  glands,  178. 

Bushing  for  link  hanger,  1(16. 

Bushing  for  reach-rod  pin  in  lifting  shaft,  106. 

Bushing  for  reach-rod  pin  in  reverse  lever.  109. 

Bushing  for  throttle  valve-rod  gland,  355. 

Bushing  for  valve-rod  end,  79. 


Cab  brackets,  52!l. 

Cabs,  533. 

Capacity,  hauling,  of  a  locomotive  on  a  grade,  615. 

Capacity  of  air  chambers  for  pumps.  :!!l(l. 

Capacity  of  pumps.  .'!!IK. 

Capacity  of  sand-boxes,  510. 


636 


INDEX. 


Capacity  of  smoke-boxes,  473. 

Capacity  of  tanks  for  tenders,  567. 

Cap  for  engine-frame  pedestals,  number  of  bolts  in,  194. 

Caps,  cast-iron,  on  sides  of  extension  fronts,  474. 

Caps  for  engine-frame  pedestals,  thickness  of,  194. 

Caps  for  frame  pedestals,  182. 

Caps  on  main  crank-pin  brasses,  316. 

Cards,  ideal  indicator,  597. 

Cards,  taken  with  an  indicator,  606. 

Case-hardening  of  crank-pins,  317. 

Casing  for  dome,  373. 

Casing  on  throttle  lever-rod,  brass,  355. 

Casting  for  front  end  of  frame  splice,  200. 

Cast-iron  ash-pan,  494. 

Cast-iron  guides,  168. 

Cast-iron  links,  102. 

Cast-iron,  selection  of,  for  cylinders,  23. 

Cause  of  variable  motion  of  crank-pin,  50. 

Cellar,  oil,  for  Pennsylvania  R.  R.  driving  axle-box,  414. 

Center-bearing  tender-trucks,  564. 

Center  line  of  motion  of  valve  gear,  43. 

Center  lines,  position  of,  on  frames,  190. 

Center  of  eccentric,  43. 

Center  of  gravity  of  a  body,  224. 

Center  of  gravity  of  an  arc,  257. 

Center  of  gravity  of  any  two  bodies,  240. 

Center  of  gravity  of  counterbalance,  229. 

Center  of  gravity  of  crank,  to  avoid  finding,  247. 

Center  of  gravity  of  engines,  position  of,  408. 

Center  of  gravity  of  lead  in  rim  of  wheel,  257. 

Center  of  gravity  of  safety-valve  levers,  377. 

Center  of  saddle-pin,  to  find  position  of,  123. 

Center  of  travel  of  valve,  important  position,  44. 

Centers  of  axles,  greatest  distance  between,  188. 

Central  side-rod  for  consolidation  engines,  289. 

Central  side-rod  for  consolidation  engines,  cross-section- 
al area  of,  312. 

Chamber  in  steam-gauge  stand  for  valves,  cocks,  etc., 
366. 

Chamber  on  top  of  boiler  for  receiving  valves,  etc.,  561. 

Change  in  appearance  of  locomotives,  3. 

Change  of  lead  with  shifting  link,  93. 

Changing  position  of  eccentric-rods,  result  of,  122. 

Check  valves,  391. 

Chime  whistle,  383. 

Cinder-box,  manner  of  opening  the,  475. 

Cinder-box  on  extension  fronts,  474. 

Circular  arc  on  throttle  valve-rod  jaw,  length  of,  363. 

Classification  of  boilers,  421. 

Classification  of  driving  axle-boxes,  203. 

Classification  of  links,  99. 

Classification  of  locomotives,  3. 

Classification  of  pistons,  138. 

Classifications  of  slide  valves,  39. 

Clearance  between  driving-boxes  and  wheels,  203. 

Clearance  between  ends  of  link  and  link  block,  130. 

Clearance  between  equalizing  lever  and  boiler  pads,  402. 

Clearance  between  wheels  of  pony  truck  and  cylinder 
heads,  542. 

Clearance,  effect  on  expansion,  603. 

Clearance  of  axle-boxes  in  pedestals,  vertical,  189. 

Clearance  of  slide  valve,  59. 

Clearance  of  slide  valve,  its  purpose,  59. 

Clearance,  piston  and  engine,  23. 

Clearance,  piston,  testing  the  amount  of,  128. 

Clearance,  practical  method  of  determining  engine,  603. 

Clearances  between  flanges  of  wheels  and  rails,  261. 

Clearance,  to  indicate  on  card,  603. 

Coal,  units  of  heat  in,  424. 

Cocks,  blow-off,  in  boilers,  418. 

Cocks,  feed,  392. 

Common  center  of  gravity  of  bodies,  principles  relating 
to,  239. 

Common  safety  valves,  371. 

Comparison  between  old  and  new  locomotives,  2. 

Comparison  of  strength  of  riveted  joint  with  that  of 
solid  plate,  457. 

Complete  locomotive  valve  gear,  35. 


Complicating  influence  of  the  angularity  of  the  eccentric- 
rod,  36. 

Compound  engines,  always  working  as  compounds,  620. 

Compound  engines,  appearance  of,  620. 

Compound  engines,  compression  of  steam  in,  618. 

Compound  engines,  cylinder  condensation  in,  617. 

Compound  engines,  diameter  of  cylinders  for  two-cylin- 
der, 628. 

Compound  engines,  diameter  of  cylinders  for  four-cylin- 
der, 631. 

Compound  engines,  engine  clearance  in,  618. 

Compound  engines,  expansive  working  of  steam  in,  619. 

Compound  engines,  experiments  with,  620. 

Compound  engines,  four-cylinder,  629. 

Compound  engines,  fuel  saved  in,  620. 

Compound  engines  in  railroad  service,  632. 

Compound  engines,  inside  clearance  of  valves  in,  618. 

Compound  engines,  intercepting  valves,  A.  J.  Pitkin,  620. 

Compound  engines,  intercepting  valve,  C.  H.  Batchellor, 
625. 

Compound  engines,  piston  valve  for  four-cylinder,  629. 

Compound  engines,  pressure  on  crank-pins  in,  619. 

Compound  engines,  ratio  of  expansion  in,  619. 

Compound  engines,  receiver  in,  620. 

Compound  engines,  re-evaporation  in  cylinders  in,  618. 

Compound  engines,  steam-port  area  in,  627. 

Compound  engines,  two-cylinder,  620. 

Compound  engines,  valve-gear  adjustment  in  two-cylin- 
der, 627. 

Compound  engines,  variation  of  pressure  in  cylinders  in, 
617. 

Compound  engines,  volume  of  receiver  in,  629. 

Compound  engines,  working  as  compounds  or  non-com- 
pounds, 625. 

Compound  locomotives,  617. 

Compression  line,  607. 

Compression  of  helical  springs,  386. 

Compression,  to  find  point  of,  60. 

Compressive  force  on  main-rods,  301. 

Computation  of  breadth  and  thickness  of  link,  105. 

Computation  of  breadth  of  bumper  beams,  520. 

Computation  of  common  center  of  gravity  of  three  equal 
weights,  241. 

Computation  of  common  center  of  gravity  of  four  equal 
weights,  241. 

Computation  of  common  center  of  gravity  of  five  equal 
weights,  242. 

Computation  of  depth  of  equalizing  lever  in  four-wheeled 
engine  trucks,  538. 

Computation  of  diameter  of  cylinder,  12,  595. 

Computation  of  diameter  of  rocker  shaft,  75. 

Computation  of  mean  effective  pressure  witli  aid  of 
ordinates,  607. 

Computation  of  mean  effective  pressure  with  aid  of  plani- 
meter,  607. 

Computation  of  pressure  at  any  point  of  stroke,  598. 

Computation  of  required  mean  effective  steam  pressure, 
596. 

Computation  of  steam-pipe  area,  30. 

Computation  of  steam-pipe  area  with  aid  of  table,  31. 

Computation  of  steam-port  area,  27. 

Computation  of  strength  of  crown  bars,  466. 

Computation  of  terminal  pressure,  599. 

Computation  of  travel  of  valve,  40. 

Computation  of  thickness  of  cylinder  walls,  25. 

Computation  of  thrust  against  guides,  159. 

Computation  of  tractive  force,  14,  16.  17,  595. 

Computation  of  weight  on  driving  wheels,  595. 

Computation  of  weight  on  engine  trucks,  15. 

Computation  of  width  of  rocker-arms,  77. 

Computations  relating  to  the  safety  valve,  374. 

Condition  and  xise  of  sand,  511. 

Conditions  a  slide  valve  must  fulfill,  38. 

Conditions  under  which  side-rods  work,  305. 

Connecting-rod,  practical  example  in  finding  thrust  of, 
158. 

Connecting-rod,  to  find  thrust  of,  156. 

Connection  of  eccentric-rods  to  link  affects  the  lead.  129. 


INDEX. 


637 


Connection  of  eccentric-rods  to  links,  88,  95. 
Connection  of  reach-rod  for  link  motion,  106. 
Connections  for  throttle  valvi',  :!.">'_'. 
C'oniifftions  for  throttle  valve  with  rod  through  side  of 

dome,  liGO. 

Connections  of  T-  and  dry-pipes,  349. 
Consolidation  engine  built  by  the  Baldwin  Locomotive 

Works,  589. 
Consolidation  engine,  computation  of  weight  on  trucks, 

15. 

Consolidation  engine,  driving-wheel  spring-gear  for,  408. 
Consolidation  engine,  frame  braces  in  a,  .~>J4. 

Consolidation  engine,  frame  for.  lit"). 

Consolidation  engine,  main-rod  for.  2H6. 

Consolidation  engines,  diameter  side-rod  bolts,  298. 

Consolidation  engines,  distance  between  centers  of  axles, 
188. 

Consolidation  engines,  driving  axle-boxes  for,  212. 

Consolidation  engines,  front  splice  for,  20d,  201. 

Consolidation  engines,  knuckle  joint  in  side-rods  for,  339. 

Consolidation  engines,  pedestals  for,  L'lL". 

Consolidation  engines,  side-rods  for.  312. 

Consolidation  engines,  side-rod  straps,  294. 

Consolidation  engines,  size  of  driving  axle-journals,  210. 

Consolidation  engines,  thickness  of  side-roil  straps,  295. 

Consolidation   locomotives,  diameter  of  driving  wheels 
of,  10. 

Consolidation  locomotives,  table  of  weights  and  hauling 
capacity  of,  19. 

Constant  lead  with  stationary  link,  88,  89. 

Construction  of  Allen  valve,  68. 

Construction  of  bells,  514. 

( 'onstruction  of  extension  fronts,  472. 

Construction  of  lifting  shaft,  106. 

Construction  of  rectangular  hyperbola,  599. 

Construction  of  reverse  lever,  108. 

Construction  of  sand-boxes,  511. 

Construction  of  slide  valve,  practical,  57. 

Construction  of  smoke-box  shell,  475. 

Construction  of  throttle  pipe,  344. 

Copper  ferrules  on  boiler  tnlies,  438. 

Copper  fire-boxes,  462. 

Copper  strips  between  guides  and  blocks,  152. 

Correct  motion  of  valve  will  depend  on  position  of  saddle- 
pin,  lifting  shaft,  and  length  of  arms,  118. 

Cotton  or  woolen  waste  in  tender-truck  axle-boxes,  578. 

Counterbalance,  amount  of.  22S. 

Counterbalance,  area  of  surface  of,  249. 

Counterbalance,  bolts  for  segments  of,  242. 

Counterbalance,  center  of  gravity  of,  229. 

Counterbalance,  common  center  of  gravity  of  two  seg- 
ments, 236. 

Counterbalance,  common  center  of  gravity  of  three  seg- 
ments, 238. 

Counterbalance,  common  center  of  gravity  of  four  seg- 
ments, 238. 

Counterbalance,   common   center  of  gravity  of  five  seg- 
ments, 239. 

Counterbalance,  cubic  inches  in  two  segments  of  a,  237. 

Counterbalanced,  weights  to  lie.  243. 

Counterbalance  for  an  eight-wheeled  engine,  243. 

Counterbalance  for  lifting  shaft.  107. 

Counterbalance,  form  of  segment  of,  -1— 

Counterbalance  for  Mogul,   ten-wheeled,  and  consolida- 
tion engines.  2">0. 

Counterbalance,  geometrical  method  of  finding  the  cen- 
ter of  gravity  of.  2311. 

Counterbalance  in  driving  wheels.  221. 

Counterbalance  in  rim  of  wheel,  effect  of  lead.  L'.">7. 

Counterbalance  in  rim  of  wheel,  lend.  2~>.~i. 

Counterbalance  in  two  segments,  to  find  weight  of  each, 
236. 

Counterbalance,  lead,  2">n. 

Counterbalance,  lever  arm  of.  242. 

Counterbalance,  number  of  cubic  inches  in  a,  23:1. 

Counterbalance,  number  of  segments  in.  I'll. 

Counterbalance,    practical   method   of  finding  the  center 
of  gravity  of,  231. 


Counterbalance,  practical  method  of  finding  the  common 
center  of  gravity  of  a  number  of  segments,  239. 

Counterbalance,  thickness  of,  235. 

Counterbalance,  thickness  of  solid,  248. 

Counterbalance,  thickness  of  two  segments  in  a,  237. 

Counterbalance,  to  compute  the  weight  of  a  simple,  225. 

Counterbalance,  to  find  weight  of,  to  balance  crank  and 
an  additional  load  on  crank-pin,  233. 

Counterbalance,  use  of  two  weights  or  segments  in,  235. 

Counterbalance,  weight  of  lead  in  rim  of  wheel,  255. 

Counterbalance,  weight  of  solid,  249. 

Counterbore,  23. 

Counter-pressure  line,  607. 

Courses  in  extension  fronts,  number  of,  475. 

Cover  for  tender-truck  axle-boxes,  578. 

Crank-pin  brasses  for  solid-ended  side-rods,  326. 

Crank-pin,  eight-wheeled  engine,  dimensions  of  main-rod 
journal  on,  319. 

Crank-pin  for  eight-wheeled  engine,  main,  319. 

Crank-pin  hub,  weight  of,  258, 

Crank-pin,  length  of  journal  of,  318. 

Crank-pin,  main,  side-rod  journal  on,  for  eight-wheeled 
engines,  320. 

Crank-pin,  ratio  of  length  and  diameter,  318. 

Crank-pin,  relation  of  its  motion  to  that  of  piston,  49. 

Crank-pins,  317. 

Crank-pins,  collars  on,  323. 

Crank-pins  for  eight-wheeled  engines,  318. 

Crank-pins  for  eight-wheeled  engines,  diameter  of  wheel 
fit,  325. 

Crank-pins-  for   Mogul,  ten-wheeled,  and  consolidation 
engines,  328. 

Crank-pins  for  Mogul,  ten-wheeled,  and  consolidation  en- 
gines, wheel  fit,  337. 

Crank-pins  for  narrow-gauge  engines,  332. 

Crank-pins  for  side-rods  for  consolidation  engines,  335. 

Crank-pins  for  side-rods  for  Mogul  and  ten-wheeled  en- 
gines, 332. 

Crank-pins  for  solid-ended  side-rods,  325,  326. 

Crank-pins,  heating  of,  318. 

Crank-pin,  side-rod   journal   on   main,  for   Mogul,  ten- 
wheeled,  and  consolidation  engines,  329. 

Crank-pins,  pressure  for  forcing  into  wheel,  326. 

Crank-pins,  pressure  required  to  force  into  hub,  260,  318. 

Crank-pins,  projected  area  of,  318. 

Crank-pins,  shank  of,  323. 

Crank-pins,  strength  of,  318. 

Crank-pin,  to  find  position  of,  for  full  and  half  stroke,  120. 

Crank,  relative  position  to  that  of  eccentric,  46. 

Crank,  to  find  dead  centers  of,  97. 

Crank,  to  find  center  of  gravity  of,  234. 

Crank,  to  find  position  of  piston  corresponding  to  that 
of,  51. 

Crank,  weight  of,  referred  to  the  crank-pin,  228. 

Crosby  pop  safety  valves,  385. 

Crosshead,  built-up,  152. 

Crosshead  classified,  149. 

Crosshead,  clearance  between  guide  blocks  and,  166. 

Crosshead  gibs,  149,  151. 

Crosshead  gibs,  area  of  sliding  surface  of,  163. 

Crosshead  gibs,  length  and  breadth  of,  164. 

Crosshead  gibs,  thickness  of,  Ki'i. 

Crosshead  gib,  wear  of  upper,  Ui'J. 

Crosshead,  greatest  pressure  of,  against  guides,  163. 

Crosshead  guides,  149. 

Crosshead  hubs,  diameter  of,  165. 

Cni-shcad  hubs,  stress  in,    Mifi. 

Crosshcad  keys,  149. 

Crosshead.  manner  of  fastening  to  piston-rod,  145. 

Crosshead-pin.  149. 

Crosshead-pin,  position  of,  153. 

Crosshead-pin.  diameter  of.  174. 
Crosshead-pin,  length  of.  173. 
Crosshcail-]>in.  pressure  on,  174. 
Crosshead-pin.  projected  area  of,   174. 
Crosshead-pins.  dowel-pins  for,  1  70. 
Crosshetid-pins.  form  of.   17o. 
Crosshcad-piiiK,  heating  of,  174. 


638 


INDEX. 


Crosshead-pina,  proportions  of,  172. 

Crosshead-pins,  reamers  for,  176. 

Crosshead-pins,  securing  of,  176. 

Crosshead-pins,  taper  of,  176. 

Crosshead-pin,  strength  of,  174. 

Crosshead,  position  of  cylinders  will  determine  the  form 
of,  150. 

Crosshead,  pressure  between  guides  and,  153,  163. 

Crossheads,  function  of,  149. 

Crossheads,  proportions  of,  163. 

Crossheads  with  babbitted  wings,  151. 

Crossheads  with  glass  disks  in  wings,  151. 

Crossheads,  width  of,  165. 

Crosshead  wings,  149. 

Crosshead  with  four  slides,  149,  150. 

Crosshead  with  four  slides,  loose  pin,  152. 

Crosshead  with  one  guide,  155. 

Crosshead  without  gibs,  case-hardened  guides.  151. 

Crosshead  with  pin  cast  in  one  piece,  objections  to,  151. 

Cross-sectional  area  of  exhaust  pipes,  507. 

Cross-sectional  area  of  frame  braces,  192. 

Cross-sectional  area  of  frame  brace,  slab  form,  197. 

Cross-sectional  area  of  full-stroke  pump-plunger,  396. 

Cross-sectional  area  of  rectangular  part  of  throttle  pipe, 
349. 

Cross-sectional  area  of  rim  of  wheel,  256. 

Cross-sectional  area  of  side-rods,  308. 

Cross-sectional  area  of  stacks,  502. 

Cross-section  of  exhaust  passage,  32. 

Crown-bar  bolts,  465. 

Crown-bar  braces,  466. 

Crown-bar  braces,  cross-sectional  area  of,  467. 

Crown  bars,  464. 

Crown  bars,  distance  between  crown  sheet  and,  464. 

Crown  bars,  position  of,  464. 

Crown-bar  thimbles,  465. 

Crown-bar  washers.  465. 

Crown  sheet,  form  of,  432,  463. 

Crown  sheet,  form  of,  in  Belpaire  boilers,  463. 

Crown  sheet,  radial  stay  bolts  to,  463. 

Crown  sheet,  sloping,  430. 

Curvature  of  reversing  links,  difference  in,  88. 

Curvature  of  throttle  lever  quadrants,  to  determine,  357. 

Curves  on  railroads,  degree  of,  614. 

Curves  on  railroads,  radius  of,  614. 

Curves,  resistance  due  to.  614. 

Cut-off  affected  by  lead,  53. 

Cut-off  affected  by  length  of  eccentric-rods,  73. 

Cut-off  affected  by  travel  of  valve,  54. 

Cut-off,  equalized,  72. 

Cut-off,  equalized,  affected  by  position  of  lifting  shaft,  73. 

Cut-off  equalized  by  means  of  link,  72. 

Cut-off,  equalized,  exhaust  regular  with,  73. 

Cut-off,  point  of,  39. 

Cut-offs,  adjusted  with  links,  89. 

Cut-offs,  sharp,  obtained  with  a  long  travel  of  valve,  40. 

Cut-off,  to  find,  length  of  connecting-rod  given,  69. 

Cut-off,  to  find  point  of,  52,  60. 

Cutting  off  early,  advantage  of,  G09. 

Cylinder  and  cylinder  heads,  joints  between,  23. 

Cylinder  bridges,  22. 

Cylinder  cocks,  22. 

Cylinder  condensation  and  re-evaporation,  617. 

Cylinder  condensation  cannot  be  detected  on  indicator- 
cards,  618. 

Cylinder  flanges,  thickness  of,  24. 

Cylinder-head  bolts,  objection  to,  24. 

Cylinder-head  bolts,  stress  in,  24. 

Cylinder-head  bolts,  to  find  diameter  of,  23. 

Cylinder  heads,  21. 

Cylinder  heads,  thickness  of,  24. 

Cylinder-head  studs,  24. 

Cylinder  lagging,  24. 

Cylinder  lagging,  asbestos  paper  for,  24. 

Cylinder  oil-cups,  554. 

Cylinders,  21. 

Cylinder  saddles  cast  separate,  20. 

Cylinders,  arrangement  of  steam-ways  and  ports,  20. 


Cylinders  cast  with  half  saddles,  20. 

Cylinders,  classification  of,  20. 

Cylinders,  computation  of  diameter,  1 2,  595. 

Cylinders,  manner  of  fastening  to  frame,  20. 

Cylinders,  position  of,  American  practice,  20. 

Cylinders,  selection  of  iron  for,  23. 

Cylinder,  thickness  of  bridges,  25. 

Cylinder  walls,  computation  of  thickness  of,  25. 

Cylinder  walls,  thickness  of,  24. 

Cylinder  walls,  thickness  of,  in  ferry  boats,  25. 


Data  required  for  choice  of  type  of  locomotive,  5. 

Data  used  for  computing  weight  of  locomotives,  15. 

Dead  bars  in  water  grates,  489. 

Dead  centers  of  crank,  to  find,  97. 

Dead  centers,  to  locate,  126. 

Degree  of  curves  on  railroads,  614. 

Dents  in  side  of  fire-box,  437. 

Depth  of  brass  packing-rings  for  pistons,  144. 

Depth  of  central  side-rods  for  consolidation  engines, 
312. 

Depth  of  engine-frame  pedestals,  189. 

Depth  of  equalizing  lever  for  pony  trucks,  546. 

Depth  of  equalizing  lever  in  four-wheeled  engine  trucks, 
538. 

Depth  of  equalizing  levers  for  driving-wheel  springs, 
402,  405. 

Depth  of  fires  in  locomotive  boilers,  424. 

Depth  of  flanges  of  tires,  263. 

Depth  of  frame  splices  for  Mogul  and  consolidation  en- 
gines, 201. 

Depth  of  front  frame  splice,  199. 

Depth  of  grate  bars  for  burning  wood,  484. 

Depth  of  lower  engine  frame  brace,  193. 

Depth  of  main-rods,  301,  303. 

Depth  of  piston  at  center,  139. 

Depth  of  piston  spider,  139. 

Depth  of  side-rods  for  eight-wheeled  engines,  309. 

Depth  of  side-rods  for  Mogul  engines,  310. 

Depth  of  single  guides,  167. 

Depth  of  upper  frame  brace,  193. 

Depth  of  upper  frame  brace,  slab  form,  197. 

Depths  of  hard  and  soft  coal  burning  furnaces,  428. 

Designing  driving  axle-boxes,  206. 

Designing  engine  frames,  188. 

Design  of  front  end  of  main-rod,  267. 

Design  of  reverse  lever,  108. 

Design  of  tender-truck  axle-boxes,  578. 

Diagram  showing  the  events  of  the  distribution  of  steam, 
63. 

Diameter  and  depth  of  mud  drums,  417. 

Diameter  and  design  of  boiler,  443. 

Diameter  of  boiler  tubes,  437. 

Diameter  of  bolts  for  segments  of  counterbalance,  242. 

Diameter  of  bolts  in  main-rod,  267. 

Diameter  of  bolts  in  side-rods,  267. 

Diameter  of  bolts  through  frame  and  splice,  198. 

Diameter  of  crosshead-hubs,  165. 

Diameter  of  crosshead-pins,  174. 

Diameter  of  cylinder-head  bolts,  to  find,  23. 

Diameter  of  cylinders,  computation  of,  12,  595. 

Diameter  of  cylinders  for  two-cylinder  compounds,  628. 

Diameter  of  cylinders  for  four-cylinder  compounds,  631. 

Diameter  of  driving-axle  journal,  209. 

Diameter  of  driving  wheels,  limit  of,  9. 

Diameter  of  driving  wheels,  to  compute,  9. 

Diameter  of  eccentric,  81. 

Diameter  of  eccentric,  to  determine  the,  82. 

Diameter  of  exhaust  nozzle  orifice,  507. 

Diameter  of  front  and  rear  side-rod  pin  journals  for 
Mogul  and  ten-wheeled  engines,  333. 

Diameter  of  full-stroke  pump  plungers,  398. 

Diameter  of  guide  bolts,  168. 

Diameter  of  holes  in  tube  sheet  for  copper  ferrules,  438. 

Diameter  of  link  saddle-pin,  106. 

Diameter  of  main  driving-axle  journal,  212. 


IXDEX. 


639 


Diameter  of  main-rod  crank-pin  journal  for  Mogul,  ten- 
whcelcd,  and  cotisolidat  ion  engines,  328. 

Diameter  of  piston-mils,  145,  147. 

Diameter  of  pop  satVly  valves,  :!st. 

Diameter  of  reverse-lever  pin,  109. 

Diameter  of  rivets  in  boilers,  450. 

Diameter  of  rocker  shaft,  74. 

Diameter  of  short-stroke  pump  plunders,  398. 

Diameter  of  side-rod  bolts  for  consolidation  engines,  298. 

Diameter  of  side-rod  bolts  for  eight-wheeled  engines,  297. 

Diameter  of  side-rod  bolts  for  Mogul  engines,  2!ts. 

Diameter  of  side-rod  bolts  for  ten-wheeled  engines,  298. 

Diameter  of  side-rod  pin  journals  for  consolidation  en- 
gines, 336. 

Diameter  of  smallest  hole  that  can  be  punched  through 
boiler  plates,  449. 

Diameter  of  stacks,  502. 

Diameter  of  studs  for  stuffing-boxes,  180. 

Diameter  of  throttle  pipes,  :!4!>. 

Diameter  of  throttle  valves,  349. 

Diameter  of  tubes  affects  heating  surface,  441. 

Diameter  of  wedge  bolts,  186. 

Diameter  of  wheel  centers,  standard,  217. 

Diameters  of  wheel  fits  of  crank-pins,  325,  337. 

Diamond  stack  with  double  shell,  496. 

Diamond  stack  with  single  shell.  4!Mi. 

Diaphragm  plates  in  extension  fronts,  474. 

Diaphragm  plates  in  extension  fronts,  position  of,  474. 

Difference  between  tender  and  engine  truck  axles,  582. 

Difference  in  curvature  of  reversing  links,  88. 

Difference  in  lead  is  affected  by  the  length  of  eccentric- 
rods,  94. 

Difficulty  with  skeleton  links,  103. 

Dimensions  of  crosshead-pins,  173. 

Dimensions  of  elliptical  springs  for  lifting-shaft  counter- 
balance, INK. 

Dimensions  of  engine  frame  braces,  191. 

Dimensions  of  engine  frame  pedestals,  186. 

Dimensions  of  guides,  166. 

Dimensions  of  lifting  shaft  and  its  arms,  107. 

Dimensions  of  link-blocks.  101. 

Dimensions  of  link  hanger,  106. 

Dimensions  of  main-rod  crank-pin  journal  for  eight- 
wheeled  engines,  .'119. 

Dimensions  of  pistons  and  rods,  146. 

Dimensions  of  reach-rods  for  link  motion,  108. 

Dimensions  of  rockers,  78. 

Dimensions  of  rod  brasses,  317. 

Dimensions  of  tender-truck  axle  journals,  581. 

Dimensions  of  tender-truck  axle  journals,  to  compute, 
582. 

Dimensions  of  volute  springs  for  lifting-shaft  counter- 
balance, 108. 

Dip  pipe  in  an  air  chamber  of  pumps,  390. 

Dished  driving  wheels.  220. 

Distance  between  axles  in  consolidation  engines,  188. 

Distance  be'twcen  axles  in  eight-wheeled  engines,  188. 

Distance  between  axles  in  Mogul  engines,  188. 

Distance  between  axles  in  ten-wheeled  engines,  188. 

Distance  between  backs  of  flanges  of  tires.  2o'<!. 

Distance  between  bulls  in  cylinder  heads.  23. 

Distance  between  bottom  of  crown  bars  and  crown  sheet, 
464. 

Distance  between  centers  of  axles,  greatest,  188. 

Distance  between  centers  of  boiler  stay  bolts,  463. 

Distance  between  centers  <>f  engine-truck  axles,  538. 

Distance  between  crown  bars.  464. 

Distance  bet  ween  eccentric-rod  pin  arc  and  link  arc,  104. 

Distance  between  guides,  four  in  a  set,  153. 

Distance  between  guides,  two  in  a  set.   l.'i.'i,  170. 

Distance  between  pony  truck  wheels  and  cylinder  heads, 
542, 

Distance  between  throttle-lever  handle  and  reverse  lever, 
356. 

Distance  from  center  of  axle  to  fire-box,  82. 

Distance  from  center  of  safety  valve  to  fulcrum,  377, 
380. 

Distance  of  rivets  from  edge  of  plate,  448. 


Distribution  of  flanged  tires  in  various  classes  of  loco- 
motives, 261. 

Distribution  of  steam  by  slide  valve,  38. 

Distribution  of  steam,  events  of,  60. 

Dome  casing,  373. 

Domes  on  boilers,  470. 

Dome  top,  joint  between  ring  and,  381. 

Dome  top  made  in  two  pieces,  387. 

Dome-top  ring,  381. 

Dome  tops,  380. 

Door  sheet,  inclination  of  furnace,  432. 

Doors  in  front  of  smoke-boxes,  476. 

Double  exhaust  pipes,  504. 

Double  exhaust  pipes,  advantages  of,  506. 

Double  poppet  throttle  valve,  343. 

Double-riveted  lap  joint,  piteh  of  rivets  in,  455. 

Double-riveted  lap  joints,  454. 

Dowel  pins  for  crosshead-pins,  176. 

Draft  pipe  in  short  smoke-boxes,  470. 

Draw  bars,  529. 

Draw-heads  for  tender,  565. 

Draw-head,  wrought-iron,  on  bumper  beam,  517. 

Driving  axle-box  brasses,  204. 

Driving  axle-box  brasses,  pressure  for  forcing  into  box, 
205. 

Driving  axle-boxes,  203. 

Driving  axle-boxes,  classification  of,  203. 

Driving  axle-boxes  for  Mogul  and  consolidation  engines, 
212. 

Driving  axle-boxes,  play  between  wheels  and,  203. 

Driving  axle-boxes,  proportions  of,  205. 

Driving  axle-box  oil-cellars,  204. 

Driving  axle-box,  Pennsylvania  R.  R.,  413. 

Driving  axle-box,  Pennsylvania  R.  R.,  cellar,  414. 

Driving  axle-box,  pockets  in,  204. 

Driving  axle-box,  width  of,  208. 

Driving  axle  journal,  length  of,  208. 

Driving   axle   journals   for   eight-wheeled,    Mogul,  ten- 
wheeled,  and  consolidation  engines,  210. 

Driving  axles,  214. 

Driving  axle,  wheel  fits  on,  214. 

Driving-boxes,  location  of  center  lines  on  engine  frames, 
190. 

Driving-boxes,  vertical  clearance  in  pedestals  for,  189. 

Driving-boxes,  vertical  movement  of,  189. 

Driving  wheel,  center  of  gravity  of  lead  in  rim  of,  257. 

Driving-wheel  counterbalance,  221. 

Driving-wheel  covers,  530. 

Driving  wheel,  cross-sectional  area  of  rim  of,  256. 

Driving  wheel,  effect  of  lead  counterbalance  in  the  rim 
of,  257. 

Driving-wheel  keys,  220. 

Driving  wheels,  215. 

Driving  wheels,  clearance  between  rails  and  flanges  of, 
261. 

Driving  wheels,  diameter  of,  9,  10. 

Driving  wheels,  fitting  up  a  pair  of,  260. 

Driving  wheels,  form  of  tread  of,  263. 

Driving  wheels,  pressure'  for  forcing  on  axle,  220. 

Driving-wheel    spring   gear   for   eight-wheeled    engines, 

400. 
!  Driving-wheel  spring  gear  for  ten-wheeled  engines,  400. 

Driving-wheel  springs,  414. 

Driving-wheel  springs,  deflection  of,  416. 

Driving-wheel  springs,  length  of,  414. 
\  Driving-wheel  springs,  load  on.  402. 

Driving-wheel  springs,  number  of  blades  in,  415. 

Driving-wheel  springs,  set  of.  41."). 

I  >ri\  ing-wheel  springs,  span  of,  414. 

Driving-wheel  springs,  thickness  of  blade's  of.  415. 

Driving  wheels,  standard  form  of  tread,  265. 

Driving  wheels,  tables  of  diameters  of,  10. 

Driving  wheels,  taper  of  tread  of.  263. 

Driving  wheels,  to  compute  diameter  of.  !l. 

Driving  wheels,  to  compute-  weight  em,  .r>9!i. 

Driving  wheels,  to  fine!  number  of.  s. 

Driving  wheels,  to  find  weight  on,  8. 

Driving  wheels  to  suit  a  particular  service,  9. 


640 


INDEX. 


Driving  wheels,  weight  of  rail  must  be  known  to  deter- 
mine the  number  of,  8. 

Driving  wheels,  weight  on,  limited  by  bridges,  8. 
Driving  wheels  with  solid  spokes,  217. 
Driving  wheels  with  tires  bolted  to  them,  260. 
Driving-wheels,  215. 
Driving-wheel  tires,  259. 
Drop  plate  in  grates,  486. 
Drop-plate  shaft,  486. 

Dry  pipe  and  throttle  pipe,  ball  joint  between,  347. 
Dry  pipe  and  T-pipe,  ball  joint  between,  350. 
Dry-pipe  and  T-pipe  connections,  349. 
Dry  pipe  for  throttle  valve,  344,  346. 
Dry-pipe  ring  on  front  flue  sheet,  ball  joint  between,  347. 
Dry  pipes,  diameter  of,  346,  349. 
Dry  pipes  on  crown  bars,  348,  373. 
Dry  pipes,  rivets  through  sleeve  and,  346. 
Dry  pipes,  suspension  of,  349. 
Dry-pipe  yoke,  347. 
Dunbar  piston  packing,  142. 
Dust-guard  in  tender-truck  axle-boxes,  578. 
Duty  of  eccentrics,  80. 
Duty  of  exhaust  passage,  22. 
Duty  of  slide  valves,  35. 
Duty  of  steam  passage,  22. 
Duty  of  steam-pipes,  30. 
Duty  of  steam-ways,  22. 


Early  cut-offs,  advantage  of,  609. 

Eccentric,  action  of,  80. 

Eccentric,  angular  advance  of,  45. 

Eccentric,  angular  advance  of,  for  shifting  links,  88,  90. 

Eccentric,  angular  advance  of,  for  stationary  links,  88,  90. 

Eccentric,  angular  advance  of,  when  rocker  is  used,  112. 

Eccentric,  backward,  89. 

Eccentric,  center  of,  43. 

Eccentric,  diameter  of,  81. 

Eccentric,  forward,  89. 

Eccentricity  of  eccentric,  37,  82. 

Eccentricity  of  eccentric,  relation  of  throw  to,  41. 

Eccentricity  of  eccentric,  to  find,  83. 

Eccentric,  key-ways  in,  81. 

Eccentric,  line  from  which  it  is  set,  43. 

Eccentric,  position  of,  rocker-arms  of  equal  lengths,  111. 

Eccentric,  position  of  (rocker  not  used),  41. 

Eccentric,  position  of  (valve  with  lead),  43. 

Eccentric,  relative  position  to  that  of  crank,  46. 

Eccentric-rod,  backward,  89. 

Eccentric-rod,  complicating  influence  of,  36. 

Eccentric-rod,  definition  of  length  of,  89. 

Eccentric-rod,  forward,  89. 

Eccentric-rod  pin  arc,  100. 

Eccentric-rod  pin  arc,  distance  between  link  arc  and,  104. 

Eccentric-rod  pin  bushing,  103. 

Eccentric-rod  pin  bushing,  loose,  103. 

Eccentric-rod  pin  bushing,  pressure  required,  103. 

Eccentric-rod  pin  bushing,  thickness  of,  103. 

Eccentric-rod  pins,  distance  between  the,  103. 

Eccentric-rod  pins,  size  of,  103. 

Eccentric-rod  pins,  wear  of,  1(13. 

Eccentric-rods,  connection  of,  to  links,  88. 

Eccentric-rods,  connection  to  link  affects  the  lead,  129. 

Eccentric-rods,  correct  way  of  connecting  to  links,  95. 

Eccentric-rods,  introduction  of  link  will  change  lengths 
of,  118. 

Eccentric-rods,  length  of,  affects  difference  in  lead,  94. 

Eccentric-rods,  length  of,  affects  the  cut-off,  73. 

Eccentric-rods,  long,  tend  to  give  a  symmetrical  motion 
to  the  valve,  37. 

Eccentric-rods,  manner  of  fastening,  to  eccentric,  81. 

Eccentric-rods  of  infinite  length  give  symmetrical  mo- 
tion, 37. 

Kccentric-rods,  result  of  changing  posit imi  of,  122,  129. 

Eccentric-rods,  to  connect  to  link,  129. 

Eccentric-rods,  to  find  correct  lengths  of,  122. 

Eccentrics,  amount  of  throw  of,  104. 

Eccentrics  and  straps,  79. 


Eccentrics,  duty  of,  80. 

Eccentrics  made  in  one  or  two  parts,  81. 

Eccentrics  not  on  main-axle,  valve  motion  with,  136. 

Eccentrics,  position  of,  for  full  and  halt'  stroke,  123. 

Eccentrics,  position  of,  rocker-arms  of  unequal  lengths, 

114. 
Eccentrics,  position  of,  rocker  employed,  center  line  of 

motion  of  valve  gear   not   coinciding  with   that  of 

piston,  116. 

Eccentrics,  practical  way  of  setting,  47. 
Eccentrics,  proportions  of,  84. 
Eccentrics,  reason  for  adopting,  38,  80. 
Eccentrics,  set  screws  in,  8] . 
Eccentrics,  space  for,  82. 
Eccentric-strap,  form  of,  81. 
Eccentric-strap  joints,  81. 
Eccentric,  throw  of,  37,  82. 
Eccentric,  to  determine  diameter  of,  82. 
Eccentric,  to  find  angular  advance  of,  97,  99. 
Eccentric,  to   find   angular   advance   of,  for   stationary 

links,  91. 

Eccentric,  to  find  position  of,  43. 
Eccentrics,  to  find  position  of,  for  full  and  half  stroke, 

121. 

Eccentrics,  to  find  position  of,  on  axle,  132. 
Eccentric  will  travel  ahead  of  crank,  when  no  rocker  is 

used,  42. 

Effect  of  lap  on  linear  and  angular  advance,  45. 
Effect  of  lead  counterbalance  in  rim  of  wheel,  257. 
Effect  of  tension  on  riveted  joints,  448. 
Efficiency  of  single-riveted  lap  joints,  458. 
Eight-wheeled   engine   built  by  the   Grant  Locomotive 

Works,  585. 
Eight-wheeled  engine  built  by  the  Pennsylvania  R.  R., 

585. 

Eight-wheeled  engine,  counterbalance  for  an,  243. 
Eight-wheeled  engine,  driving-wheel  spring  gear  for,  400. 
Eight-wheeled  engine,  frame  braces  in  an,  523. 
Eight-wheeled  engines,  crank-pins  for,  318. 
Eight-wheeled  engines,  cross-sectional  area  of  side-rod 

bolts  for,  297. 
Eight-wheeled  engines,  cross-sectional  area  of  side-rods, 

308. 
Eight-wheeled  engines,  diameter  of  side-rod  bolts  for, 

297. 

Eight-wheeled  engines,  distance  between  axles,  188. 
Eight-wheeled  engines,  driving  axle-boxes  for,  210. 
Eight-wheeled  engines,  driving  axle-journals  for,  210. 
Eight-wheeled  engines,  main  crank-pin  for,  319. 
Eight-wheeled  engines,  position  of  side-rods  in,  290. 
Eight-wheeled  engine,  spring  gear  for  narrow-gauge,  403. 
Eight-wheeled  engines,  side-rods  for  narrow-gauge,  290. 
Eight-wheeled  engines,  stress  in  side-rod  bolts  for,  297. 
Eight-wheeled  engines,  stress  in  side-rod  straps,  288. 
Eight-wheeled  engines,  thickness  and  depth  of  side-rods, 

309. 
Eight-wheeled  engines,  thickness  of  side-rod  straps  for. 

286. 
Eight-wheeled  locomotives,  computation   of   weight   on 

trucks,  15. 
Eight-wheeled  locomotives,  diameter  of  driving  wheels, 

10. 
Eight-wheeled  locomotives,  table  of  weight  and  hauling 

capacity,  17. 

Elevated  Railroad,  smoke-boxes  on  the  New  York,  471. 
Elevated  roads,  main-  and  side-rods  for  engines  on,  278. 
Elliptical  springs  for  lifting-shaft  counterbalance,  107. 
Elliptical  springs  for  lifting-shaft  counterbalance,  dimen- 
sions of,  108. 

Encased  pop  safety  valves,  385. 
Energy  and  work,  612. 

Energy  required  to  overcome  the  force  of  gravity,  612. 
Engine  clearance,  23. 

Engine  clearance  and  piston  displacement,  ratio  of,  603. 
Engine  clearance,  effect  on  expansion,  603. 
Engine  clearance  in  compounds,  618. 
Engine  clearance,  practical  method  of  determining,  603. 
Engine  clearance,  to  indicate  on  card,  603. 


641 


Engine  draw-head,  front,  517. 

Engine,  finish  at  front  end  of,  522. 

Engine  frame  ami  splice,  diameter  of  bolts  through,  198. 

Engine  frame  and  splice  forged  in  one  piece,  201. 

Engine-frame  bolts,  201. 

Engine  frame  braces,  dimensions  of,  191. 

Engine-frame  braces,  equal  depth  throughout,  196. 

Engine-frame  brace,  slab  form,  cross-sectional  area  of, 

HIT. 

Engine-frame  braces,  s I ress  per  square  inch,  191. 

Engine  frame,  built-up,  I!*."). 

Engine  frame  for  a  consolidation  locomotive,  195. 

Engine  frame,  light,  196. 

Engine  frame,  lower  brace,  depth  of,  193. 

Engine-frame  pedestal  legs,  thickness  of,  193. 

Engine-frame  pedestals,  1M2. 

Engine-frame  pedestals,  depth  of,  189. 

Engine  frame,  position  of  center  lines  on,  190. 

Engine  frames.  I  Ml. 

Engine  frames,  tiaek  ends  suitable  for  foot-plates,  187. 

Engine  frames,  designing,  188. 

Engine  frames  for  eight-wheeled  passenger  locomotives, 
187. 

Engine-frame  splice,  casting  for  front  end  of,  200. 

Engine-frame  splice  for  consolidation  and  Mogul  locomo- 
tives, 200,  201. 

Engine-frame  splice,  form  of  front  end  of,  200. 

Engine-frame  splice,  depth  of,  199. 

Engine-frame  splices,  recess  for  cylinder  saddle  in,  199. 

Engine  frames,  proportions  of,  187. 

Engine  frames,  slab,  196. 

Engine  frame,  upper  brace,  depth  of,  193. 

Engine-frame  wedges,  thickness  of,  186. 

Engine  frame,  width  of,  191. 

Engine  pedestal  caps,  number  of  bolts  in,  194. 

Engine  pedestal  caps,  thickness  of,  194. 

Engine  pedestals  for  Mogul,  ten-wheeled,  and  consolida- 
tion engines,  212. 

Engine  pedestals,  position  of  straight  legs,  190. 

Engine  pedestals,  taper  of  legs,  190. 

Engine  pedestals,  width  of  opening  of,  190. 

Engines,  hard-coal  burning,  424. 

Engines,  soft-coal  burning,  421. 

Engines,  space  required  between  axles  in  soft-coal  or 
wood-burning,  188. 

Engine-truck  axle-boxes,  548. 

Engine-truck  axle-journals,  549. 

Engine-truck  axle-journals,  diameter  and  length  of,  550. 

Engine-truck  axle-journals,  pressure  on,  549. 

Engine-truck  center-plates,  538. 

Engine  truck,  distance  between  centers  of  axles  in,  538. 

Engine  truck,  equalizing  lever  for  four-wheeled,  538. 

Engine  truck,  four-wheeled,  535. 

Engine  truck,  length  of  two-wheeled,  542. 

Engine  truck,  Pennsylvania  K.  B.,  538. 

Engine-truck  swing-center  castings,  538. 

Engine  truck,  two-wheeled,  540. 

Engine  truck,  two-wheeled,  cast-iron  frame  for,  542. 

Engine  truck,  two- wheeled,  clearance  between  wheels 
and  cylinder  heads,  542. 

Engine  truck,  two-wheeled,  computation  of  length  of,  544. 

Engine  truck,  two-wheeled,  equalizing  lever  for.  54(i. 

Engine  truck,  two-wheeled,  graphical  method  of  finding 
length  of,  542. 

Engine  truck,  two-wheeled,  thickness  and  depth  of  equal- 
izing lever.  547. 

Engine-truck  wheel  covers,  535. 

Equalized  cut-off.  "L". 

Equalized  cut-off  affected  by  position  of  lifting  shaft,  73. 

Equalizing  lever  for  driving-wheel  springs,  depth  of, 
402.  4115. 

Equalizing  lever  for  driving-wheel  springs,  load  on,  40.'!. 

Equalizing  levers,  clearance  between  boiler  pads  and, 
402. 

Equalizing  level's  for  driving-wheel  springs,  4ou. 

Equalizing  levers  for  driving-wheel  springs,  thickness  <>(. 
402,  40*. 

Equalizing  levers,  gibs  and  plates  on  top  of,  411. 


Equalizing  levers,  purpose  of,  400. 

Equalizing  lever,  thickness  of  metal  outside  of  slots 
through,  406. 

Equalizing  the  cut-off  by  means  of  a  link,  72. 

Erecting  shop,  practical  example  of  setting  the  valve 
gear  in  a,  125. 

Evaporation,  rate  of,  on  inclined  sheets,  432. 

Events  of  the  distribution  of  steam,  60. 

Events  of  the  distribution  of  steam,  diagram  showing 
the,  63. 

Exhaust  closure,  to  find  the  point  of,  619. 

Exhaust  line,  607. 

Exhaust  nozzles,  area  of  orifice  in,  507. 

Exhaust  nozzles,  position  of,  in  long  and  short  smoke- 
boxes,  477. 

Exhaust  passage,  22. 

Exhaust  passage,  cross-section  of,  32. 

Exhaust  passage,  duty  of,  22. 

Exhaust  passage  openings,  size  of,  32. 

Exhaust  passages,  small  space  for,  33. 

Exhaust  passage,  thickness  of  metal  around,  25. 

Exhaust  pipes,  advantages  of  single  and  double,  506. 

Exhaust  pipes  and  nozzles,  504. 

Exhaust  pipes,  cross-sectional  area  of,  507. 

Exhaust  pipes,  form  of  passages  in,  506. 

Exhaust  pipes,  long,  507. 

Exhaust  pipes,  long  and  short,  where  used,  506. 

Exhaust  pipes,  short  double,  504. 

Exhaust  pipes,  short  single,  505. 

Exhaust  port,  22. 

Exhaust  ports,  area  of,  26. 

Exhaust  ports,  length  and  breadth  of,  26,  29. 

Exhaust  regular  with  equalized  cut-offs,  73. 

Exhaust  steam,  action  of,  502. 

Exhaust  thimbles,  505. 

Expansion  curve,  597,  607. 

Expansion  of  steam,  lap  required  for,  49. 

Expansion,  ratio  of,  clearance  neglected,  600. 

Expansion,  ratio  of,  with  clearance,  604. 

Expansive  working  of  steam  in  a  compound,  619. 

Experiments  with  compound  engines,  620. 

Extension  fronts,  472. 

Extension  fronts  affect  the  exhaust,  473. 

Extension  fronts,  blowing  out  by  steam,  474. 

Extension  fronts,  capacity  of,  473. 

Extension  fronts,  cast-iron  caps  on  side  of,  474. 

Extension  fronts,  cinder-box  on,  474. 

Extension  fronts,  construction  of,  473. 

Extension  fronts,  diaphragm  plates  in,  474. 

Extension  fronts,  length  of,  475. 

Extension  fronts,  long  exhaust  pipe  in,  476. 

Extension  fronts,  number  of  courses  in,  475. 

Extension  fronts,  position  of  diaphragm  plates  in,  474. 

Extension  fronts,  rings  in,  475. 


Eac-tor  of  safety  for  rod  bolts,  275. 

Fastening  domes  to  boilers,  470. 

Fastenings  for  boiler  to  frames,  482. 

Fastening  water  grates  in  furnace,  489. 

Favorite  form  of  main-rod,  301. 

Feed-cock  plug,  opening  in,  394. 

Feed-cock  quadrant,  394. 

Feed-cock  roil.  394. 

Feed-cocks,  391. 

Feed-pipe  hangers  and  connections.  :i!>4. 

Ferrule  for  reach-rod  pin  in  lifting  shaft,  106. 

Ferrule  for  reach-rod  pin  in  reverse  lever,  109. 

Ferrules,  copper,  on  boiler  tubes,  4.'i8. 

Ferry-boats  used  in  connection  with  railroads,  25. 

Finger  grate-liars.  484. 

Finish  of  front  end  of  engine.  522. 

Fire-box,  Helpaire.  432. 

Fire-box,  dents  in  sides  of,  4.'i7. 

Fire-box,  depths  of  hard-  and  soft -coal  burning,  US. 

Fire-box  designed  by  Mr.  .1.  lleadden,   1J4. 

Fire-box,  distance  from  center  of  axle  to,  82. 


642 


IXDEX. 


Fire-box,  door  and  side  sheets,  inclination  of,  432. 

Fire-boxes,  steel,  iron,  and  copper,  462. 

Fire-box,  greatest  width  of,  427. 

Fire-box,  limit  of  size,  422. 

Fire-door  opening,  436. 

Fire-box,  position  of,  for  soft-coal  or  wood-burning  en- 
gines, 188. 

Fire-box  relative  to  axles,  position  of,  421. 

Fire-box  ring,  434. 

Fire-box,  rivets  in,  435. 

Fire-box  sheets,  thickness  of,  434. 

Fire-box,  wood-burning,  430. 

Fitting  up  a  pair  of  driving  wheels,  260. 

Flanged  tires,  distribution  of,  261. 

Flanged  tires,  thickness  and  width  of,  265. 

Flanges  of  tires,  depth  of,  263. 

Flanges  of  tires,  distance  between  backs  of,  266. 

Flanges  of  wheels,  clearance  between  rails  and,  261. 

Flanges  on  driving  axle-boxes,  form  of,  204. 

Flanges  on  engine-frame  wedges,  thickness  of,  186. 

Flanges  on  rod  brasses,  316. 

Flanges,  thickness  of  cylinder,  24. 

Flexibility  of  side-rods,  306. 

Follower  bolts  for  pistons,  139. 

Follower  plate  for  pistons,  138,  139. 

Foot-plate  braces,  524. 

Foot-plates,  527. 

Foot-plates,  back  ends  of  engine  frames  suitable  for,  187. 

Force  acting  on  the  ends  of  equalizing  lever  for  pony 
truck,  546. 

Force  a  locomotive  must  exert  to  haul  a  train,  7. 

Force,  moment  of  a,  226. 

Force  of  gravity,  energy  required  to  overcome,  612. 

Force  which  a  rocker  must  overcome,  75. 

Form  and  size  of  steam  and  exhaust  ports,  26. 

Form  of  crosshead  determined  by  position  of  cylinders, 
150. 

Form  of  crosshead-pins,  176. 

Form  of  crown  sheet,  432,  463. 

Form  of  eccentric  strap,  81. 

Form  of  main-rod,  301. 

Form  of  passages  in  exhaust  pipes,  506. 

Form  of  piston  and  valve-rod  glands,  179. 

Form  of  piston  packing,  140. 

Form  of  saddles  for  smoke-stacks,  502. 

Form  of  safety-valve  seat,  371. 

Form  of  segment  of  counterbalance,  242. 

Form  of  slide-valve,  ordinary,  34. 

Form  of  templet  for  segment,  246. 

Form  of  tread  of  driving  wheels,  263. 

Form  of  tread  of  driving  wheel,  standard,  265. 

Forms  of  driving-wheel  spring  hangers,  various,  411. 

Formula  for  computing  load  on  crown  bars,  466. 

Formulas  for  computing  boiler  pressure,  462. 

Formulas  for  computing  thickness  of  boiler  plates,  462. 

Formulas,  rules,  and  data,  595. 

Forward  eccentric,  89. 

Forward  eccentric-rod,  89. 

Forward  stroke,  89. 

Four-cylinder  compound  engines,  629. 

Four-cylinder  compound  engines,  diameter  of  cylinders, 
631. 

Four-wheeled  engine  truck,  535. 

Frame  and  splice,  diameter  of  bolts  through,  198. 

Frame  bolts,  201. 

Frame  brace,  engine,  depth  of  lower,  193. 

Frame  brace,  engine,  depth  of  upper,  193. 

Frame  brace,  engine,  equal  depth  throughout,  196. 

Frame  braces,  belly  brace,  523. 

Frame  braces,  bumper  brace,  522. 

Frame  braces,  engine,  dimensions  of,  191. 

Frame  braces,  engine,  stress  per  square  inch,  191. 

Frame  braces,  guide  yoke,  522. 

Frame  braces  in  a  consolidation  engine,  524. 

Frame  braces  in  a  Mogul  engine,  524. 

Frame  braces  in  an  eight-wheeled  engine,  523. 

Frame  brace,  slab  form,  depth  of,  197. 

Frame,  engine,  built-up,  195. 


Frame,  engine,  width  of,  191. 

Frame  for  a  consolidation  engine,  195. 

Frame  for  two-wheeled  engine  truck,  542. 

Frame  pedestal  caps,  number  of  bolts  in,  194. 

Frame  pedestal  caps,  thickness  of,  194. 

Frame  pedestal,  dimensions  of,  186. 

Frame  pedestal  legs,  thickness  of,  193. 

Frame  pedestals,  182. 

Frame  pedestals,  wedges  for,  184. 

Frames,  engine,  light,  196. 

Frames  for  engine,  designing,  188. 

Frames  for  engine,  proportions  of,  187. 

Frames  for  tender,  iron,  569. 

Frames  for  tender,  wooden,  565. 

Frame  splice  and  frame  forged  in  one  piece,  201. 

Frame  splice,  casting  for  front  end  of,  200. 

Frame  splice,  depth  of,  199. 

Frame  splice  for  consolidation  and  Mogul  engines,  200. 

Frame  splices  for  passenger  locomotives,  187,  197. 

Frame  splice,  recess  for  cylinder  saddle  in,  199. 

Frames,  position  of  center  lines  on  engine,  190. 

Frames,  slab,  196. 

Freight  locomotive,  management  of,  2. 

Friction  in  the  joints  of  boiler  plates,  451. 

Friction  of  slide-valve,  64. 

Friction,  resistance  due  to,  611. 

Front  draw-head  on  engine,  517. 

Front  draw-heads  on  tenders,  567. 

Front  end  of  main-rod,  design  of,  267. 

Front  flue-sheet  ring  for  dry  pipe,  347. 

Front  side-rod  for  consolidation  engines,  289. 

Front  side-rod  for  Mogul  engines,  281. 

Front  side-rod  pins'  for  Mogul  and  ten-wheeled  engines, 

projected  area  of,  333. 

Front  side-rods  for  consolidation  engines,  312. 
Fronts  on  smoke-boxes,  476. 
Front  driving-wheel  springs  in  Mogul  and  consolidation 

engines,  load  on,  546. 

Front  strap  of  main-rod,  diameter  of  bolts  through,  279. 
Fuel  saved  in  compounds,  620. 
Fulcrum  for  driving-wheel  spring  equalizing  lever,  400, 

409. 
Fulcrum  for  driving-wheel  spring  equalizing  lever,  area 

of,  409. 
Fulcrum  for  driving-wheel  spring  equalizing  lever,  load 

on,  403. 
Fulcrum  for  driving-wheel  spring  equalizing  lever,  ratio 

of  thickness  and  width,  410. 
Fulcrum  for  driving-wheel  spring  equalizing  lever,  stress 

in,  409. 

Full  gear  of  link  motion,  89. 
Full-stroke  pump  plungers,  diameter  of,  398. 
Full-stroke  pump  plungers,  cross-sectional  area  of,  396. 
Full-stroke  pumps,  387. 
Function  of  crossheads,  149. 
Function  of  frame  pedestals,  182. 
Furrowing  in  boiler  plates,  459. 
Fusible  plugs,  435. 


Gauge  cocks,  561. 

Gear  for  throttle  valves  with  rod  through  end  of  boiler, 

352. 
Gear  for  throttle  valves  with  rod  through  side  of  dome, 

359. 
Gear  for  throttle  valves  with  rod  through  top  of  boiler, 

359,  364,  366. 
Geometrical  method  of  finding  the  center  of  gravity  of 

counterbalance,  230. 
Gib  of  crosshead,  wear  of  upper,  162. 
Gibs  and  plates  on  top  of  equalizing  lever,  411. 
Gibs  for  crossheads,  149,  151. 

Gibs  for  crossheads,  area  of  sliding  surface  of,  163. 
Gibs  for  crossheads,  length  and  breadth  of,  164. 
Gibs  for  crossheads,  thickness  of,  165. 
Gland  for  throttle  valve-rod,  355. 
Glands  and  stuffing-boxes,  proportions  of,  176. 
Glands  and  stuffing-boxes,  studs  for,  180. 


1XDEX. 


643 


Gland.-,  luisliiiig  lor,  17H. 

Gltiiids  (or  piston-rods,  'Jl. 

Glands  for  valve-roils  HIII!  piston-rods,  metal  in,  178. 

Glands,  length  of,  179. 

Gluss  disks  in  wings  of  crossheads,  151. 

Goose-neck,  393. 

Grade,  6(19. 

(trade.  hauling  capacity  of  a  locomotive  on  a,  615. 

Grade,  resistance  due  to,  612. 

i ; rate-bars,  484. 

Grate-bars,  finger,  484. 

Grate-bars  for  burning  wood,  484. 

Grate-bars  for  burning  wood,  depth  of,  484. 

i  irate-bars,  rocking,  for  burning  soft  coal,  484. 

<  irate-bars,  shaking  lever  for,  486. 

Grate  dead-bars,  489. 

i  irate,  drop-plate  in  furnace,  486. 

Orate  surface,  421. 

(irate  surface,  area  of,  soft-coal,  422. 

Grates,  water,  486. 

Gravity,  energy  required  to  overcome  the  force  of,  612. 

Guide-blocks,  149. 

Guide-blocks,  clearance  between  crosshead  and,  166. 

Guide-bolts,  168. 

Guides,  allowance  for  wear  and  re-planing,  167. 

Guides,  cast-iron,  168. 

Guides,  computation  of  thrust  against,  159. 

Guides,  copper  slips  between  blocks  and,  152. 

Guides,  depth  of  single,  167. 

Guides,  dimensions  of,  166. 

Guides  for  crossheads,  149. 

Guides,  four  in  a  set,  distance  between,  153. 

Guides,  greatest  pressure  of  crosshead  against,  163. 

Guides,  length  and  breadth,  166. 

Guides,  practical  examples  of  finding  thrust  against,  158. 

Guides,  pressure  between  crosshead  and,  153,  163. 

Guides,  tapered,  167. 

Guides,  thrust  against,  156. 

Guides,  to  compute  thickness  of,  166. 

Guides,  two  in  a  set,  distance  between,  155,  170. 

Guide-yoke,  149,  522. 

Gusset  plates  in  boiler,  470. 


Half  gear  of  link  motion,  89. 

Hand  holes  in  boilers,  417. 

Handle  for  opening  safety-valves,  362,  369. 

Handle  for  shaking  lever,  486. 

Hanger  for  link,  100. 

Hanger  for  link,  dimensions  of,  106. 

Hanger,  link,  87. 

Hangers  for  driving-wheel  springs,  400,  410. 

Hangers  for  driving-wheel  springs,  cross-sectional  area 
of,  410. 

Hangers  for  driving-wheel  springs,  ratio  of  thickness  to 
width,  411. 

Hangers  for  driving-wheel  springs,  stress  per  square  inch, 
410. 

Hangers  for  driving-wheel  springs,  tension  on,  403. 

Hangers  for  driving-wheel  springs,  various  forms  of,  411. 

Hangers  for  feed-pipe,  .'i'J4. 

Hard-coal  burners,  position  of  axles  under,  189. 

Hauling  capacity  of  a  locomotive  on  a  grade,  615. 

Hauling  capacity  of  locomotives,  table  of,  17,  18,  19. 

Heads  for  pistons,  138. 

Heads,  thickness  of  cylinder,  24. 

Heater  pipe,  394. 

Heating  of  crank-pins,  318. 

Heating  of  crosshead-pins,  174. 

Heating  of  driving-box  brasses,  206. 

Heating  surface  affected  by  diameter  of  tube,  441. 

Heating  surface  per  cubic  foot  of  piston-drop  displace- 
ment, 445. 

Heating  surface  proportioned  to  tractive  power,  446. 

Heating  surface,  ratio  of  fire-box  to  tube,  446. 

Height  of  stacks,  503. 

Height  of  valve-seats,  29. 


Helical  springs,  compression  of,  386. 

Helical  springs,  size  of  steel  for,  386. 

Hole  in  top  of  balanced  slide-valves,  67. 

Holes  in  rocker-arms,  taper  of,  74. 

Holes  in  tube  sheet  for  copper  ferrules  on  tubes,  diam- 
eter of,  438. 

Hole,  the  smallest  that  can  be  punched  through  boiler 
plates,  449. 

Hollow  brick  arches,  477. 

Hollow  stay  bolts  in  boilers,  463. 

Horse -power,  605. 

Hose,  suction,  393. 

House  brackets,  529. 

Houses,  533. 

Hub  on  rocker-arms,  78. 

Hubs  for  crossheads,  diameter  of,  165. 

Hubs  for  crossheads,  stress  in,  165. 

Hyperbola,  rectangular,  598. 

Hyperbola,  rectangular,  construction  of,  599. 

Hyperbolic  logarithms,  601. 


I  cross-section  for  side-rods,  307. 

Important  position,  center  of  the  travel  of  valve  is  an,  44. 

Inclination  of  furnace  door  and  side  sheets,  432. 

Inclination  of  water  grates,  486. 

Indicator-card,  ideal,  597. 

Indicator-card,  scale  for,  603. 

Indicator-cards,  measurement  of  pressures  on,  598. 

Indicator-cards  taken  with  an  indicator,  606. 

Indicator,  steam  accounted  for  by  the,  607. 

Infinite  lengths  of  eccentric-rods  give  symmetrical  mo- 
tion, 37. 

Influence  of  engine  clearance  on  expansion,  603. 

Influence  of  rocker-arms  and  link  on  travel  of  valve,  37. 

Influence  of  rocker  on  the  motion  of  the  valve,  36. 

Initial  steam  pressure,  597. 

Injectors,  558. 

Injector  stop-valve,  561. 

Inside  clearance,  its  purpose,  59. 

Inside  clearance  of  valves  in  compounds,  618. 

Inside  lap,  59. 

Inside  lead,  59. 

Intercepting  valve  in  compounds,  A.  J.  Pitkin,  620. 

Intercepting  valve  in  compounds,  C.  H.  Batchellor,  625. 

Introduction  of  link  will  change  the  lengths  of  eccentric- 
rods,  118. 

Introduction  of  link  will  not  change  position  of  eccentrics 
nor  rockers,  118. 

Iron  fire-boxes,  462. 

Iron  pilot,  518. 

Iron  plates,  tensile  strength  of,  451. 

Iron  tender-frame,  569. 

Irregular  surfaces,  area  of,  254. 


Jaw  for  throttle-valve  rod,  353,  362. 

Jaw  for  throttle-valve  rod,  length  of  circular  arc  on,  363. 

Joint  between  dome-top  and  ring,  381. 

Joint  between  throttle  and  dry  pipes,  ball,  347. 

Joint  in  valve-rod,  knuckle,  76,  79. 

Joints  between  cylinder  and  cylinder  heads,  23. 

Joints  for  steam-pipes,  ball,  .'13. 

Joints  in  boiler  plates,  friction  in,  451. 

Joints  in  boilers,  butt,  460. 

Joints  in  eccentric-straps,  81. 

Joints,  pitch  of  rivets  in  double-riveted  lap,  4.V>. 

Joints,  pitch  of  rivets  in  single-riveted  lap,  450,  453. 

Journal,  diameter  of  driving-axle,  209. 

Journal,  length  of  driving-axle,  208. 

Journal  of  main  crank-pin  for  Mogul,  ten-wheeled,  mid 

consolidation  engines,  328. 
Journal  of  main  driving-axle,  diameter  of,  212. 
Journal   of  side-rod    journals  for  consolidation  engines, 

336. 

Journals  for  tender-truck  axles,  580. 
Journals  for  tender-truck  axles,  diameter  of,  581. 


644 


INDEX. 


Journals  for  tendei'-truck  axles,  pressure  on,  581. 

Journals  of  crank-pin,  eight-wheeled  engines,  318. 

Journals  of  driving-axle,  pressure  per  square  inch  of  pro- 
jected area,  206. 

Journals  of  driving-axle,  projected  area  of,  206. 

Journals  of  front  and  rear  side-rod  pin  for  Mogul  and 
ten-wheeled  engines,  333. 

Journals  on  engine-truck  axles,  549. 

Journals  on  engine-truck  axles,  diameter  and  length  of. 
550. 

Journals  on  engine-truck  axles,  pressure  on,  549. 

Journals,  to  find  diameter  and  length  of,  when  the  ratio 
between  them  is  known,  551. 


Keys  for  crossheads,  149. 

Keys  for  cylinder  saddle  and  front  splice,  199. 

Keys  for  driving-wheels,  220. 

Keys  for  main-  and  side-rods,  form  of,  299. 

Keys  for  main-  and  side-rods,  thickness,  width,  and  taper 

of,  300. 

Keys  for  piston-rods,  taper  of,  145. 
Keys  for  pistons,  strength  of,  148. 
Keys  in  valve-rod  end,  79. 
Keys,  taper  of  main-  and  side-rods,  270. 
Key,  transverse,  in  front  end  of  main-rod,  300. 
Key-way  in  axle  for  eccentrics,  81. 
Key-ways  in  eccentric,  81. 
Knuckle-joint  for  side-rods,  338. 
Knuckle-joint  in  valve-rod,  79. 
Knuckle-joint  of  side-rods,  pin  through,  281. 
Knuckle-joint  pin  for  side-rods,  pressure  on,  340. 


Lagging,  asbestos  paper  for  cylinder,  24. 

Lagging  on  boiler,  extension  of,  374. 

Lagging  on  boiler,  thickness  of,  366. 

Lamp  bracket  for  steam-gauge,  362. 

Lap  and  travel  of  valve,  to  find,  56. 

Lap  at  each  end  of  valve  is  equal,  72. 

Lap,  effect  of,  on  linear  and  angular  advance,  45. 

Lap,  inside,  59. 

Lap  joints,  double-riveted,  454. 

Lap  joints,  pitch  of  rivets  in  double-riveted,  455. 

Lap  joints,  single-riveted,  447. 

Lap  joints,  single-riveted,  efficiency  of,  458. 

Lap  of  valve,  to  find,  55. 

Lap,  outside,  39. 

Lap,  problems  relating  to,  of  the  slide-valve,  52. 

Lap  required  for  the  expansion  of  steam,  49. 

Latch  for  reverse  lever,  109. 

Latch  for  reverse  lever,  position  of,  109. 

Latch,  May's  reverse  lever,  110. 

Late  cut-offs,  disadvantage  of,  609. 

Laying  out  a  valve  gear,  118. 

Lead  affected  by  the  connection  of  eccentric-rods  to  link, 

129. 

Lead  affects  point  of  cut-off,  53. 
Lead,  amount  of,  95. 
Lead,  change  of,  with  shifting  link,  93. 
Lead  constant  with  stationary  link,  88,  89. 
Lead  counterbalance,  250. 
Lead  counterbalance  in  rim  of  wheel,  255. 
Lead  counterbalance  in  rim  of  wheel,  effect  of,  257. 
Lead  in  rim  of  wheel,  center  of  gravity  of,  257. 
Lead  in  rim  of  wheel,  weight  of,  255. 
Lead  is  equal  at  each  end  of  valve,  full  stroke,  72,  102. 
Lead,  length  of  eccentric- rods  effects  difference  in,  94. 
Lead  lining  for  tender-truck  axle-box  brasses,  577. 
Lead  lining,  preparation  of  brass  for,  577. 
Lead  of  valve,  42. 

Lead  variable  with  shifting  link,  88. 
Leaves  in  driving-wheel  springs,  number  of,  415. 
Leaves  in  driving-wheel-springs,  thickness  of,  415. 
Legs,  straight,  engine  pedestal,  position  of,  190. 
Legs,  taper  of  engine  pedestal,  190. 
Length  and  breath  of  crosshead  gibs,  164. 


Length  of  arc  described  by  top  of  reverse  lever,  109. 

Length  of  bumper  beams,  518. 

Length  of  circular  arc  on  throttle-vaive  rod  jaw,  363. 

Length  of  crank-pin  journal,  318. 

Length  of  crosshead-pins,  173. 

Length  of  driving-axle  journal,  208. 

Length  of  driving-wheel  springs,  414. 

Length  of  eccentric-rod,  definition  of,  89. 

Length  of  eccentric-rods  affects  cut-off,  73. 

Length  of  eccentric-rods  affects  difference  in  lead,  94. 

Length  of  eccentric-rods,  to  find,  96,  129. 

Length  of  eccentric-rods  will  be  changed  by  the  intro- 
duction of  links,  118. 

Length  of  engine  pony-truck,  computation  of,  544. 

Length  of  engine-pony  truck,  graphical  method  of  find- 
ing, 542. 

Length  of  equalizing  lever  for  engine  pony-truck,  546. 

Length  of  exhaust  ports,  29. 

Length  of  front  and  rear  side-rod  pin  journals  for  Mogul 
and  ten-wheeled  engines,  333. 

Length  of  guides,  166. 

Length  of  link,  104. 

Length  of  main-rod  crank-pin  journal  for  eight-wheeled 
engines,  320. 

Length  of  main-rod  crank-pin  journal  for  Mogul,  ten- 
wheeled,  and  consolidation  engines,  328. 

Length  of  radius  for  stationary  links,  90. 

Length  of  radius  of  shifting  link,  104. 

Length  of  reverse  lever,  109. 

Length  of  rocker-arms,  76. 

Length  of  rocker-shaft,  74. 

Length  of  safety-valve  lever,  374,  376,  380. 

Length  of  side-rod  pin  journals  for  consolidation  engines, 
336. 

Length  of  side-rods,  how  taken,  317. 

Length  of  smoke-stacks,  503. 

Length  of  steam-ports,  26. 

Length  of  throttle-lever  quadrants,  to  determine  the,  357. 

Length  of  tubes,  442. 

Length  of  valve-seat,  30. 

Lengths  of  eccentric-rods,  to  find  correct,  122. 

Lever  arm  of  counterbalance,  242. 

Lever  arm,  or  arm  of  a  force,  225. 

Lever  for  opening  safety-valve,  362,  369. 

Lever  for  shaking  grates,  486. 

Lever  for  throttle  valve,  352,  354,  356. 

Lever  for  throttle  valve,  position  of,  355. 

Lever,  reverse,  87,  108. 

Levers  for  driving-wheel  springs,  depth  of  equalizing, 
400,  402,  405. 

Levers  for  driving-wheel  springs,  equalizing,  400. 

Levers  for  driving-wheel  springs,  thickness  of  equalizing, 
402,  408. 

Levers  for  driving-wheel  springs,  thickness  of  metal  out- 
side of  slots  through,  406. 

Levers  for  whistle,  383. 

Lifting  shaft,  87,  106. 

Lifting-shaft  arms,  87,  107. 

Lifting-shaft  arms,  dimensions  of,  107. 

Lifting-shaft  counterbalance  springs,  dimensions  of,  108. 

Lifting  shaft,  diameter  of,  107. 

Lifting  shaft,  location  of,  107. 

Lifting  shaft,  object  of,  36. 

Lifting-shaft  pin,  10(1,  106. 

Lifting  shaft,  position  for  full  gear  backward,  106. 

Lifting  shaft,  position  for  full  gear  forward,  106. 

Lifting  shaft,  position  for  midgear,  106. 

Lifting  shaft,  to  find  position  of,  124. 

Lifting  shaft  will  affect  the  cut-off,  position  of,  73 

Lift  of  pump  valves,  389. 

Lift  of  throttle  valve,  358. 

Limited  number  of  bolts  in  side-rods,  296. 

Limit  of  diameter  of  driving  wheels,  9. 

Limit  of  size  of  fire-box,  422. 

Limit  of  wear  of  tires,  265. 

Linear  advance  affected  by  lap,  45. 

Linear  advance  of  valve,  45. 

Line,  atmospheric,  602,  607. 


IXDEX. 


645 


Line  from  which  eccentric  is  sot,  43. 

Line  of  admission.  (!07. 

Line  nf  compression,  (ill". 

Line  of  counter-pressure,  (>07. 

Line  of  direction  of  a  force,  220. 

Line  of  exhaust.  b'i>7. 

Line  of  perfect  vacuum,  .">!•".  tin". 

Line  of  perfect  vacuum,  location  of,  602. 

Liners  for  main-  and  side-rods,  270. 

Lines,  position  of  center,  on  engine  frames,  190. 

Link  arc,  definition  of,  99. 

Link  arc,  distance  between  eccentric  rod-pin  arc  and, 
104. 

Link-block,  87,  101. 

Link-block  pin,  100. 

Link-block  pin,  diameter  of,  100. 

Link-blocks,  proportions  of,  101. 

Link-blocks,  slip  of,  102. 

Link,  box,  Km. 

Link,  breadth  of.  In.-,. 

Link,  clearance  between  link-block  and  ends  of,  130. 

Link,  definition  of  radius  of,  89,  100. 

Link  hanger,  87,  100,  10(5. 

Link-hanger  bushing.  KMi. 

Link  hanger,  dimensions  of,  106. 

Link,  introduction  of  link  will  not  change  position  of 
eccentrics  or  rockers,  118. 

Link,  length  of,  99. 

Link  motion,  86. 

Link  motion,  dimensions  of  reach-rod  for,  108. 

Link  motion,  reach-rod  for,  87. 

Link  motion,  reach-rod  for,  connection  of,  106. 

Link,  open,  100. 

Link,  radius  of,  to  find,  121. 

Link  saddle,  87,  100. 

Link  saddle-piu,  87,  100. 

Link  saddle-pin,  diameter  and  length  of,  106. 

Links,  cast-iron,  102. 

Links,  classification  of,  99. 

Links,  connection  of  eccentric-rods  to,  88. 

Links,  difference  in  curvature  of,  88. 

Links,  difficulty  with  skeleton,  103. 

Link,  skeleton,  99. 

Links,  method  of  attaching,  87. 

Link,  solid,  99. 

Links,  point  of  cut-off  adjusted  with,  89. 

Links,  proportions  of,  102. 

Links,  safety,  between  engine  and  tender,  528. 

Links,  shifting,  88. 

Links,  stationary,  88. 

Links,  steel,  102. 

Link,  Stephenson's,  object  of,  36. 

Link,  suspension  of,  100. 

Links,  wrought-iron,  102. 

Link,  thickness  of,  105. 

Link,  to  connect  eccentric-rods  to,  129. 

Link  used  as  a  means  for  equalizing  the  cut-off,  72. 

Load  on  driving-wheel  equalizing  lever,  403. 

Load  on  driving-wheel  springs,  402. 

Load  on  front  springs  in  ten-wheeled,  Mogul,  and  con- 
solidation engines,  546. 

Load  supported  by  crown  bars,  466. 

Locate  the  dead  centers,  to,  1L'(>. 

Location  of  reverse  lever,  108,  373. 

Location  of  rocker,  118. 

Locomotive,  consolidation,  built  by  the  Baldwin  Loco- 
motive Works,  589. 

Locomotive,  eight-wheeled,  built  by  the  Grant  Locomo- 
tive Works,  585. 

Locomotive,  eight-wheeled,  Pennsylvania  R.  R.  standard, 
BBS. 

Locomotive  frame,  built-up,  195. 

Locomotive  frame  for  a  consolidation  engine,  195. 

Locomotive  frames,  ISL'. 

Locomotive  frames,  light,  196. 

Locomotive,  hauling  capacity  on  a  grade,  615. 

Locomotive,  ten-wheeled,  built  by  the  Baldwin  Loco- 
motive Works,  589. 


Locomotive,  practical  example  of  setting  the  valve  gear 
of  a,  125. 

Locomotive  pumps,  location  of,  585. 

Locomotives,  advantage  of  present  management  of,  2. 

Locomotives,  change  in  appearance  of,  3. 

Locomotives,  classification  of,  3. 

Locomotives,  compound,  617. 

Locomotives,  comparison  between  old  and  new,  2. 

Locomotives,  compound,  built  by  the  Baldwin  Locomo- 
tive Works,  629. 

Locomotives,  compound,  built  by  the  Rhode  Island  Loco- 
motive Works,  625. 

Locomotives,  compound,  built  by  the  Schenectady  Loco- 
motive Works,  620. 

Locomotives,  consolidation,  3. 

Locomotives,  consolidation,  distance  between  centers  of 
axles,  188. 

Locomotives,  data  required  for  choice  type  of,  5. 

Locomotives,  data  used  for  computing  weight  of,  15. 

Locomotives,  diameter  of  driving  wheels  for  consolida- 
tion, 10. 

Locomotives,  diameter  of  driving  wheels  for  eight- 
wheeled,  10. 

Locomotives,  diameter  of  driving  wheels  for  Mogul,  10. 

Locomotives,  diameter  of  driving  wheels  for  ten-wheeled, 
10. 

Locomotives,  eight-wheeled,  3. 

Locomotives,  eight-wheeled,  distance  between  centers  of 
axles,  188. 

Locomotives,  force  that  must  be  exerted  to  haul  a  train, 
7. 

Locomotive,  six-wheeled  shifting,  built  by  the  Pennsyl- 
vania R.  R.,  589. 

Locomotives,  large,  limit  of  total  wheel-base,  189. 

Locomotives,  management  of,  2. 

Locomotives,  Mogul,  3. 

Locomotives,  Mogul,  distance  between  centers  of  axles, 
188. 

Locomotives,  parts  made  stronger  of  modern,  2. 

Locomotives,  points  of  support  of,  408. 

Locomotives,  position  of  center  of  gravity,  408. 

Locomotives,  small,  total  wheel-base  of,  189. 

Locomotives,  ten-wheeled,  3. 

Locomotives,  ten-wheeled,  distance  between  centers  of 
axles,  188. 

Locomotives,  types  of,  585. 

Locomotives,  weight  of,  15. 

Locomotives,  work  done  by,  12. 

Logarithms,  hyperbolic,  601. 

Long  eccentric-rods  tend  to  give  a  symmetrical  motion  to 
the  valve,  37. 

Long  exhaust  pipes,  507. 

Long  exhaust  pipes,  where  used,  506. 

Longitudinal  bolts  through  tender-frames,  563. 

Longitudinal  play  of  driving  boxes,  203. 

Longitudinal  stress  in  boiler  shell,  447. 

Long  smoke-boxes,  472. 

Long  smoke-boxes  affect  the  exhaust,  473. 

Long  smoke-boxes,  blowing  out  by  steam,  474. 

Long  smoke-boxes,  cast-iron  caps  on  side  of,  474. 

Long  smoke-boxes,  cinder-box  on,  474. 

Long  smoke-boxes,  construction  of,  47:i. 

Long  smoke-boxes,  diaphragm  plates,  474. 

Long  smoke-boxes,  length  of,  475. 

Long  smoke-boxes,  netting  in,  474. 

Long  smoke-boxes,  petticoat  pipe  in,  476. 

I. ung  smoke-boxes,  rings  in,  475. 

Long  wedge  for  frame  pedestals,  184. 

Lower  engine-frame  brace,  depth  of,  193. 


Main  crank-pin  for  a  consolidation  engine,  328. 
Main  crank-pin  for  eight-wheeled  engine's.  l!l!>. 
Main  driving-axle  journals,  diameter  of,  212. 
Main-rod,  area  of  smallest  transverse  section  of,  301. 

Main-rod  bolts.  272. 
Main-rod  brasses,  .'!!.">. 


646 


INDEX. 


Main-rod,  design  of  front  end  of,  267. 

Main-rod,  favorite  form  of,  301. 

Main-rod  for  erosshead  with  two  slides,  300. 

Main-rod  journal  on  crank-pin,  eight-wheeled  engine, 
dimensions  of,  319. 

Main-rod,  liners  for,  270. 

Main-rod,  number  and  diameters  of  bolts  in,  267. 

Main-rods  and  side-rods,  267. 

Main-rods,  diameter  of  bolts  through  front  strap  of,  279. 

Main-rods,  diameter  of  bolts  through  rear  strap  of,  276. 

Main-rods  for  a  consolidation  engine,  286. 

Main-rods  for  engines  on  elevated  roads,  278. 

Main-rods,  form  of  keys  for,  299. 

Main-rods  for  Mogul  engines,  281. 

Main-rods,  material  for,  303. 

Main-rods,  number  of  bolts  through  rear  strap  of,  276. 

Main-rods,  oil-holes  in  straps  for,  296. 

Main-rods,  pressure  per  square  inch  of  cross-section,  301. 

Main-rods,  ratio  of  depth  and  thickness  of,  301. 

Main-rods,  sections  of,  to  be  avoided,  303. 

Main-rods,  tensile  and  compressive  force  on,  301. 

Main-rods,  thickness  and  depth  of,  301,  303. 

Main-rods,  thickness,  width,  and  taper  of  keys  for,  270, 
300. 

Main-rod  straps,  stress  per  square  inch  of  cross-sectional 
area  of,  285. 

Main-rod  straps,  thickness  of,  283. 

Main-rods  with  chamfered  edges,  303. 

Main-rod  with  transverse  key  in  front  end,  300. 

Management  of  locomotives,  2. 

Manner  of  fastening  piston-rods,  145. 

Mariotte's  law,  598. 

Material  for  piston-rods,  147. 

Material  for  valve-rod  and  piston-rod  glands,  178. 

May's  reverse-lever  latch,  110. 

Mean  effective  pressure  computed  with  the  aid  of  ordi- 
nates,  607. 

Mean  effective  pressure  computed  with  the  aid  of  plani- 
meter,  607. 

Mean  effective  steam  pressure,  computation  of  required, 
596. 

Mean  effective  steam  pressure  for  a  given  cut-off,  597, 
602,  605. 

Mean  pressure,  601. 

Mean  pressure,  computation  of,  602. 

Measurement  of  pressure  on  cards,  598. 

Metal  in  rod  brasses,  317. 

Method,  correct,  of  connecting  eccentric-rods  to  links,  95. 

Method  for  determining  engine  clearance,  practical,  603. 

Method  of  attaching  links,  87. 

Method  of  equalizing  cut-off,  73. 

Method  of  finding  the  amount  of  rolling  and  axle  fric- 
tion, 6. 

Method  of  finding  the  tractive  power  or  force,  11. 

Method  of  setting  a  simple  valve-gear,  practical,  96. 

Mid-gear  of  link  motion,  89. 

Mogul  engine,  counterbalance  for  a,  250. 

Mogul  engine,  driving-wheel  spring-gear  for  a  narrow- 
gauge,  408. 

Mogul  engine,  frame  braces  in  a,  524. 

Mogul  engine  frame  splices,  200,  201. 

Mogul  engines,  computation  of  weight  on  trucks,  15. 

Mogul  engines,  crank-pins  for,  328. 

Mogul  engines,  diameter  of  driving  wheels  of,  10. 

Mogul  engines,  diameter  of  side-rod  bolts  for.  298. 

Mogul  engines,  distance  between  centers  of  axles,  188. 

Mogul  engines,  driving  axle-boxes  for,  212. 

Mogul  engines,  knuckle-joint  in  side-rods  for,  338. 

Mogul  engines,  least  thickness  of  side-rod  straps  for,  294. 

Mogul  engines,  pedestals  for,  212. 

Mogul  engines,  side-rod  pins  for,  332. 

Mogul  engines,  side-rods  for,  281,  309. 

Mogul  engines,  si/e  of  driving-axle  journals,  210. 

Mogul  engines,  stress  in  side-rod  straps  for,  291. 

Mogul  engines,  thickness  of  straps  for  side-rods  for,  292. 

Mogul  engine,  table  of  weights  and  hauling  capacity  of, 
18. 

Moment  of  a  force,  226,  378. 


Moment  of  steam  pressure  on  safety-valves,  375. 
Moment  of  tension  on  safety-valve  springs,  375;  378. 
Motion  of  valve  influenced  by  the  rocker,  36. 
Motion  of  valve  will  depend  on  position  of  saddle-pin, 

lifting  shaft,  length  of  its  arms,  119. 
Movement  of  axle-boxes,  189. 
Mud  drums,  417. 
Mud  plugs  in  boilers,  417. 

N 

Names  of  lines  on  ideal  indicator-cards,  597. 

Narrow-gauge  eight-wheeled  engines,  side-rods  for,  290. 

Narrow-gauge  engines,  crank-pins  for,  332. 

Netting  in  extension  fronts,  474. 

New  York  Elevated  Kailroad,  smoke-boxes  on,  471. 

Normal  pressure  on  rails,  610. 

Notches  in  reverse-lever  quadrants,  109. 

Notches  in  reverse-lever  quadrant,  to  lay  off,  133. 

Nozzles  for  exhaust  pipes,  504. 

Number  of  bolts  in  a  segment  of  counterbalance,  242. 

Number  of  bolts  in  engine-frame  pedestal  caps,  194. 

Number  of  bolts  in  guides,  168. 

Number  of  bolts  in  main-rods,  267. 

Number  of  bolts  in  side-rods,  267. 

Number  of  bolts  in  side-rods,  limited,  297. 

Number  of  bolts  through  frame  and  splice,  198. 

Number  of  courses  in  smoke-box  shell,  475. 

Number  of  driving  wheels,  to  find,  8. 

Number  of  driving  wheels,  weight  of  rails  must  be  known 

to  determine,  8. 

Number  of  leaves  in  driving-wheel  springs,  415. 
Number  of  safety-valves,  362. 
Number  of  segments  in  counterbalance,  244. 
Number  of  studs  in  stuffing-boxes,  180. 
Number  of  tubes,  440. 
Nuts  on  tubes  for  supporting  brick  arches,  478. 


Objections  to  bolts  in  cylinder  heads,  24. 

Objections  to  crosshead  with  pin  cast  in  one  piece,  151. 

Objections  to  octagonal  driving-box  brasses,  204. 

Object  of  lifting  shaft,  36. 

Object  of  rocker,  36. 

Object  of  Stephenson's  link,  36. 

Oblique  braces,  stress  in,  481. 

Offset  in  rocker,  117. 

Offset  in  rocker,  to  find,  1 19. 

Oil-cellars  for  driving  axle-boxes,  204. 

Oil-cellars  for  driving  axle-box,  Pennsylvania  R.  K.,  414. 

Oil-cups,  552. 
I  Oil-cups  for  cylinders,  554. 

Oil-grooves  in  driving-box  brasses,  205. 

Oil-holes  in  main-  and  side-rod  straps,  296. 

Opening  for  fire  door,  436. 
I  Opening,  width  of  engine  pedestal,  190. 

Open  links,  99. 

Orifices  in  exhaust  nozzles,  area  of,  507. 

Outside  lap,  39. 


Pads  on  boilers,  482. 

Pads,  reinforcing  plates  inside  of  boiler  for,  483. 
Parts  to  be  counterbalanced,  243. 
Passages,  form  of  exhaust-pipe,  506. 
i  Passenger  locomotive,  management  of,  2. 
Passenger  trains,  remarks  relating  to,  2. 
Pedestal  caps,  182. 
Pedestal  jaw,  182. 

Pedestal  legs,  engine,  thickness  of,  193. 
Pedestals,  depth  of  engine-frame,  189. 
Pedestals,  engine,  vertical  clearance  for  axle-boxes,  189. 
Pedestals  for  engine  frames,  dimensions  of,  186. 
Pedestals  for  engine  frames,  function  of,  182. 
Pedestals  for  Mogul,  ten-wheeled,  and  consolidation  en- 
gines, 212. 
Pedestals,  taper  of  legs  in  engine,  190. 


IXDES. 


647 


Pedestals,  width  of  opening  of  engine,  190. 

Pennsylvania  R.  H.  engine  truck,  53H. 

Pennsylvania  H.  K.,  six-wheeled  shifting  engine  built  by 

the,  589. 

Pennsylvania  K.  R.  Standard  eight-wheeled  engine,  585. 
Perfect  vacuum  line,  f>'J7,  CUT. 
Perfect  vaeuuni  line,  location  of,  602. 

Vttieoat  pipe  in  extension  fronts,  476. 

Vtticoat  pipe  in  short  smoke-boxes,  470. 

'ilnts,  iron,  518. 

'ilots,  wooden,  517. 

'in  for  erosshead,  149. 

'in  for  lifting  shaft,  100. 

'in  for  link-block,  100. 

'in  for  link-block,  diameter  of,  100. 

'in  for  link-saddle,  87,  100. 

'in  for  reach-rod,  106. 
Pin  for  reverse-lever,  diameter  of,  109. 
Pin  in  knuckle-joint  of  side-rods,  281,  340. 
Pins  for  erossheads,  diameter  of,  174. 
Pins  for  crossheads,  length  of,  173. 
Pins  for  erossheads,  securing  of,  17(i. 
Pins  for  eccentric-rods,  distance  between,  103. 
Pins  for  eccentric-rods,  size  of,  103. 
Pins  for  lifting  shaft,  106. 
Pins  for  rockers,  74. 

Pins,  link-saddle,  diameter  and  length  of,  106. 
Pipe  connecting  throttle- valve  stuffing-box  to  dome,  361. 
Pipe  for  heating  feed-water,  394. 
Pipes,  exhaust,  504. 
Pipes,  sand,  511. 

Piston,  cause  of  its  variable  motion,  50. 
Piston  clearance,  23. 
Piston  clearance,  testing  amount  of,  128. 
Piston  displacement  and  engine  clearance,  ratio  of,  603. 
Piston  follower-bolts,  139. 
Piston  heads,  138. 

Piston,  how  to  obtain  a  symmetrical  motion,  50. 
Piston  illustrated,  146. 
Piston  keys,  139. 
Piston  keys,  strength  of,  148. 
Piston  packing,  138. 
Piston  packing,  brass,  143. 
Piston  packing,  Dunbar,  142. 
Piston  packing,  form  of,  140. 
Piston  packing  rings,  position  of,  143. 
Piston  packing,  width  of,  139. 

Piston,  relation  of  its  motion  to  that  of  crank-pin,  49. 
Piston-rod  gland,  21. 
Piston-rod  gland,  form  of,  179. 
Piston-rod  keys,  taper  of,  145. 
Piston-rods,  21. 
Piston-rods,  diameter  of,  145,  147. 

Piston-rods,  glands  for,  178. 

Piston-rods,  manner  of  fastening  to  piston,  145. 

Piston-rods,  material  for,  147. 

Piston-rods,  stress  in,  144. 

Piston-rods,  taper  at  ends,  145. 

Piston-rods,  weakest  part  of,  148. 

Pistons,  21,  138. 

Pistons,  classified,  138. 

Pistons,  depth  at  the  center,  139. 

Pistons,  follower-plate  for,  138,  139. 

Piston,  solid,  140. 

Piston  speed,  to  determine,  33. 

Piston  spider,  138. 

Piston  spider,  depth  of,  139. 

Pistons  with  one  face  spherical,  139. 

Piston,  to  find   its   position    corresponding  to  that  of 

crank,  51. 
Piston  T-rings,  140,  143. 

Piston  valve  for  four-cylinder  compounds,  020. 
Pitch  of  rivets  in  double-riveted  lap  joints,  455. 

Pitch  of  rivets  in  single-riveted  lap  joints,  453. 

Pitkin,  A.  J.,  intercepting  valves  for  compounds,  620. 

Plain  throttle-valve  lever,  352. 

Plain  tires,  261. 

Plate  for  bumpers,  522. 


Plates  and  gibs  on  top  of  equalizing  lever,  411. 

Plates,  apparent  strength  of,  451. 

Plates  for  mud  plugs  in  boilers,  reinforcing,  417. 

Plates,  friction  in  the  joints  of  boiler,  451. 

Plates  in  boilers,  formulas  for  computing  the  thickness 
of,  462. 

Plates  in  extension  fronts,  diaphragm,  474. 

Plates,  smallest  hole  that  can  be  punched  through,  449. 

Plates,  strength  of,  restored  by  annealing,  451. 

Plates,  strength  of,  restored  by  reaming  the  holes,  451. 

Plates,  stress  in  the  joints  of  boiler,  451. 

Plates,  tensile  strength  of  iron,  451,  462. 

Plates,  tensile  strength  of  steel,  462. 

Play  between  driving  boxes  and  hub  of  wheel,  203. 

Plugs,  fusible,  435. 

Pockets  in  driving  boxes,  204. 

Point  of  admission,  to  find,  60. 

Point  of  compression,  to  find,  60. 

Point  of  cut-off,  39. 

Point  of  cut-off  adjusted  with  links,  89. 

Point  of  cut-off  affected  by  lead,  53. 

Point  of  cut-off  affected  by  length  of  eccentric-rods,  73< 

Point  of  cut-off  affected  by  travel  of  valve,  54. 

Point  of  cut-off,  to  find,  52,  60. 

Point  of  cut-off,  to  find,  length  of  connecting-rod  given, 
69. 

Point  of  exhaust  closure,  to  find,  619. 

Point  of  release,  to  find,  60. 

Point  of  suppression,  to  find,  60. 

Points  of  support  of  engines,  408. 

Pony  truck,' 540. 

Pony  truck,  cast-iron  frame  for,  542. 

Pony  truck,  clearance  between  wheels  and  cylinder 
heads,  542. 

Pony  truck,  equalizing  lever  for,  546. 

Pony  truck,  graphical  method  of  finding  length  of,  542. 

Pony  truck,  length  of,  542. 

Pony  truck,  thickness  and  depth  of  equalizing  lever  for, 
547. 

Pop  safety-valve,  373,  380,  384. 

Pop  safety-valves,  Crosby,  385. 

Pop  safety-valves,  diameter  of,  384. 

Pop  safety-valves,  encased,  385. 

Pop  safety-valves,  George  W.  Richardson's,  384. 

Position  of  axle  for  laying  out  the  valve  gear,  119. 

Position  of  blow-off  cocks,  418. 

Position  of  bolts  in  side-rods  and  main-rods,  299. 

Position  of  center  lines  on  engine  frame,  190. 

Position  of  center  of  gravity  of  counterbalance,  229. 

Position  of  center  of  gravity  of  engines,  408. 

Position  of  check-valves,  391. 

Position  of  crank-pin  for  full  and  half  stroke,  to  find,  120. 

Position  of  crosshead-pin,  153. 

Position  of  crown  bars,  464. 

Position  of  cylinders,  American  practice,  20. 

Position  of  diaphragm  plates  in  smoke-boxes,  474. 

Position  of  eccentric,  line  from  which  it  is  set,  43. 

Position  of  eccentric  ;  rocker  arms  of  equal  lengths,  111. 

Position  of  eccentric :  rocker-arms  of  unequal  lengths, 
114. 

Position  of  eccentric  (rocker  not  used),  41. 

Position  of  eccentric ;  rockers  employed ;  center  line  of 
motion  of  valve-gear  not  coinciding  with  that  of  pis- 
ton, 116. 

Position  of  eccentric-rods,  result  of  changing,  1 22. 

Position  of  eccentrics  for  full  and  half  stroke,  123. 

Position  of  eccentrics  on  axle,  to  find,  132. 

Position  of  eccentric,  to  find,  44. 

Position  of  eccentric  (valve  having  lead),  43. 

Position  of  exhaust  nozzles  in  long  and  short  smoke- 
boxes,  477. 

Position  of  fire-box  for  soft-coal  burning  engines,  188. 

Position  of  fire-box  relative  to  axles,  421. 

Position  of  lifting  shaft  affects  cut-off,  73. 

Position  of  lifting  shaft,  to  find,  124. 

Position  of  notches  in  reverse-lever  quadrant,  110. 

Position  of  pump  air-chambers,  394. 

Position  of  reverse  lever,  108,  373. 


648 


IXDEX. 


Position  of  reverse-lever  latch,  109. 

Position  of  reverse  lever,  to  find,  for  full  gear  forward 
and  backward,  130. 

Position  of  rocker,  118. 

Position  of  running  boards,  532. 

Position  of  saddle-pin,  to  find,  123. 

Position  of  safety-valves,  380. 

Position  of  sand  pipes,  513. 

Position  of  side-rods  in  eight-wheeled  engines,  290. 

Position  of  steam-gauge  stand,  373. 

Position  of  straight  legs  in  engine  pedestals,  190. 

Position  of  throttle  lever,  355. 

Position  of  throttle- valve  rod  stuffing-box,  355. 

Position  of  valve,  central,  an  important  one,  44. 

Position  of  wheels  under  hard-coal  burners,  189. 

Position  of  whistle,  380. 

Position,  relative,  of  crank  to  eccentric,  46. 

Positions  of  eccentrics  for  full  and  half  stroke,  to  find, 
121. 

Power,  horse,  605. 

Power  required  to  work  a  plain  slide-valve,  63. 

Practical  construction  of  slide-valve,  57. 

Practical  example  of  finding  thrust  of  connecting-rod, 
158. 

Practical  example  of  setting  the  valve  gear  of  a  locomo- 
tive, 125. 

Practical  method  of  finding  the  center  of  gravity  of 
counterbalance,  231. 

Practical  method  of  finding  the  common  center  of  grav- 
ity of  a  number  of  segments  in  counterbalance,  239. 

Practical  method  of  setting  a  simple  valve  gear,  96. 

Practical  way  of  determining  engine  clearance,  603. 

Practical  way  of  setting  eccentrics,  47. 

Preference  of  studs  in  cylinder  heads,  24. 

Preparation  of  truck  brasses  for  lead  lining,  577. 

Pressure,  back,  602. 

Pressure  between  crossheads  and  guides,  153,  163. 

Pressure,  computation  of,  at  any  point  of  the  stroke,  598. 

Pressure,  computation  of  required  mean  effective,  596. 

Pressure,  computation  of  terminal,  599,  601. 

Pressure  for  a  given  cut-off,  mean  effective,  597,  602,  605. 

Pressure  for  eccentric-rod  pin  bushing,  103. 

Pressure  for  forcing  brasses  and  axle-boxes,  205. 

Pressure  for  forcing  bushing  into  throttle-valve  gland, 
355. 

Pressure  for  forcing  crank-pins  into  hub,  260,  326. 

Pressure  for  forcing  driving  wheels  on  axle,  220. 

Pressure,  mean,  601. 

Pressure,  mean,  computation  of,  602. 

Pressure,  mean  effective,  computed  with  the  aid  of  or- 
dinates,  607. 

Pressure,  mean  effective,  computed  with  the  aid  of  plan- 
imeter,  607. 

Pressure,  measurement,  on  cards,  598. 

Pressure  on  crank-pins,  318. 

Pressure  on  crank-pins  in  compound  locomotives,  619. 

Pressure  on  crosshead-pin,  174. 

Pressure  on  knuckle-joint  pin  in  side-rods,  340. 

Pressure  on  rails,  normal,  610. 

Pressure  on  slide-valve,  64,  65. 

Pressure  on  springs  of  pop  safety-valves,  386. 

Pressure  on  tender  and  engine-truck  axles,  difference  be- 
tween. 582. 

Pressure  per  square  inch  of  cross-section  of  main-rods, 
301. 

Pressure  per  square  inch  of  projected  area  of  driving- 
axle  journals,  206. 

Pressure  per  square  inch  of  projected  area  on  engine- 
truck  axle  journals,  549. 

Pressure  per  square  inch  on  journals  for  given  velocities, 
210. 

Pressure,  terminal,  598. 

Pressure,  variation  of,  in  cylinders,  is  reduced  in  com- 
pound engines,  617. 

Pressure  which  boilers  will  stand,  formulas  for,  462. 

Primitive  slide-valve,  48. 

Principles  relating  to  the  common  center  of  gravity  of 
bodies,  239. 


Principles  to  be  remembered  in  laying  out  a  valve  gear, 
99. 

Problems  relating  to  lap  of  slide-valve,  52. 

Problems  relating  to  slide-valves  reduced  to  simplest 
forms,  35. 

Projected  area  of  crank-pins,  318. 

Projected  area  of  crosshead-pins,  174. 

Projected  area  of  driving-axle  journals,  206. 

Projected  area  of  front  and  rear  side-rod  pins  in  Mogul 
and  ten-wheeled  engines,  333. 

Projections  on  the  inside  of  tires,  259. 

Proportions  of  crosshead-pins,  172. 

Proportions  of  crossheads,  163. 

Proportions  of  driving  boxes,  205. 

Proportions  of  eccentrics,  84. 

Proportions  of  engine  frames,  187. 

Proportions  of  link-blocks,  101. 

Proportions  of  links,  102. 

Proportions  of  main-  and  side-rod  keys,  299. 

Proportions  of  main-  and  side-rod  straps,  270. 

Proportions  of  side-rods,  308. 

Proportions  of  wedges  and  bolts  for  engine-frame  pedes- 
tals, 186. 

Provisions  for  shutting  off  steam  in  all  boiler  valves, 
cocks,  etc.,  366. 

Pulling-bar  between  engine  and  tender,  528. 

Pulling-bar  pin,  527. 

Pulling-bar  support,  528. 

Pump,  387. 

Pump  air-chambers,  390. 

Pump,  lift  of  valves,  389. 

Pump  pet-cock,  390. 

Pump  plunger,  389. 

Pump  plunger,  cross-sectional  area  of  full-stroke,  396. 

Pump  plunger,  cross-sectional  area  of  short-stroke,  398. 

Pumps,  capacity  of,  398. 

Pumps,  capacity  of  air-chambers  for,  390. 

Pumps,  dip-pipe  in  air-chamber  of,  390.. 

Pump,  size  of,  396. 

Pumps,  position  of,  394. 

Pumps,  short-stroke,  387,  395. 

Pump,  to  find  stroke  of,  398. 

Punching  reduces  tenacity  of  plates,  451. 

Punch,  spiral,  injury  to  plates,  451. 

Purpose  of  giving  lap  to  the  valve,  49. 

Purpose  of  inside  clearance,  59. 

Purpose  of  inside  lap,  59. 

a 

Quadrant  for  feed-cock,  394. 

Quadrants  for  reverse  lever,  position  of  notches  in,  110. 

Quadrants  for  reverse  levers,  109. 

Quadrants  for  throttle  lever,  to  determine  the  length  and 

curvature  of,  357. 

Quadrants  for  throttle-valve  lever,  353,  354,  355,  356. 
Quadrants,  to  lay  off  notches  on  reverse-lever,  '33. 

K 

Radial  stay-bolts  for  crown  sheet,  432,  463. 

Radius,  length  of,  for  stationary  links,  90. 

Radius  of  a  curve  on  railroads,  614. 

Radius  of  link,  definition  of,  89,  100. 

Radius  of  link,  length  of,  104. 

Radius  of  link,  to  find,  121. 

Radius  of  thi'ottle-lever  quadrant,  357. 

Railroad  service,  compound  engine  in,  632. 

Rails  can  bear,  amount  of  weight  that,  8. 

Rails,  clearance  between  flanges  of  wheels  and,  261 . 

Rails,  weight  of,  must  be  known  to  determine  number 

of  driving  wheels,  8. 
Rankine,  Prof.  W.  J.  M.,  table  of   pressures  for  given 

velocities  of  axle  journals,  210. 
Rate  of  evaporation  on  inclined  sheets,  432. 
Ratio  between  diameter  and  length  of  journal  given,  to 

find  diameter  and  length,  551. 
Ratio  of  cross-sectional  area  of  tubes  to  grate  surface, 

438. 


JXDKX. 


049 


Ratio  of  depth  and  thickness  of  main-rods,  301. 

Ratio  of  depth  uiul  thickness  of  side-rods,  309. 

Ratio  of  diameter  and  length  of  kniiekle-joint  pin,  340. 

Ratio  of  diameter  to  length  of  tubes,  441. 

Ratio  of  engine  elearanee  and  piston  displacement,  003. 

Ratio  of  expansion,  clearance  neglected,  GOO. 

Uatio  of  expansion  in  compounds,  610. 

Katio  of  expansion,  with  clearance,  604. 

Katio  of  fire-box  to  tube-heating  surface,  446. 

Katio  of  length  and  diameter  of  main  crank-pill  for  eight- 
wheeled  engines,  318. 

Kutio  of  length  and  diameter  of  main  crank-pin  journals 
for  Mogul,  ten-wheeled,  and  consolidation  engines, 
329. 

Katio  of  length  and  diameter  of  side-rod  journals  on 
main  crank-pin  for  Mogul,  ten-wheeled,  and  con- 
solidation engines,  330. 

Katio  of  thickness  to  width  of  equalizing-lever  fulcrum, 
410. 

Katio  of  thickness  to  width  of  spring  hangers,  411. 

Reach-rod  for  link  motion,  87. 

1  {each-rod  for  link  motion,  connection  of,  106. 

Reach-rod  for  link  motion,  dimensions  of,  108. 

Reach-rod  pin,  106. 

Reamers  for  crosshead-pins,  176. 

Reaming  rivet-holes  restores  strength  of  plate,  451. 

Rear  draw-heads  for  tenders,  565. 

Rear  side-rod  pins  for  Mogul  and  ten-wheeled  engines, 
projected  area  of,  333. 

Rear  side-rods  for  consolidation  engines,  289,  312. 

Rear  side-rods  for  Mogul  engines,  281. 

Rear  straps  of  main-rods,  diameter  of  bolts  through,  276. 

Receiver  in  compounds,  620. 

Receiver  in  compounds,  volume  of,  629. 

Recess  in  front  splice  for  cylinder  saddle,  199. 

Reciprocating  parts  counterbalanced,  243. 

Rectangular  hyperbola,  598. 

Rectangular  hyperbola,  construction  of,  599. 

Re-evaporation  in  compounds  is  not  wholly  lost,  618. 

Re-evaporation  in  cylinders,  617. 

Regulation  of  feed-water,  394. 

Reinforcing  plates  for  mud  plugs,  417. 

Reinforcing  plates  inside  of  boiler  for  pads,  483. 

Relation  between  motion  of  crank-pin  and  that  of  piston, 
49. 

Relative  position  of  eccentric  to  that  of  crank,  46. 

Release,  to  find  point  of,  60. 

Relief-valve  on  throttle  pipes,  345. 

Remarks  relating  to  passenger  trains,  2. 

Remarks  relating  to  valve  motions,  136. 

Re-planing  and  wear  of  guides,  allowance  for,  168. 

Resistance  due  to  curves,  614. 

Resistance  due  to  friction,  61 1. 

Resistance  due  to  grade,  612. 

Resistance  due  to  speed,  614. 

Resistance  train,  5. 

Result  of  changing  position  of  eccentric-rods,  122. 

Reverse  lever,  87. 

Reverse  lever,  construction  of,  108. 

Reverse-lever  latch,  109. 

Reverse-lever  latch,  May's,  110. 

Reverse  lever,  length  of,  109. 

Reverse  lever,  length  of  arc  described  by  top  of,  109. 

Reverse  lever,  location  of,  108,  373. 

Reverse-lever  pin,  diameter  of,  109. 

Reverse-lever  quadrants,  109. 

Reverse-lever  quadrants,  notches  in,  109. 

Reverse-lever  quadrant,  to  lay  off  notches  on,  133. 

Reverse  lever,  thickness  of,  109. 

Reverse  lever,  to  find  position  of,  for  full  gear  forward 
and  backward,  130. 

Reverse  shaft,  87. 

Reverse-shaft  arms,  87. 

Reverse-shaft  anus,  dimensions  of,  107. 

Reverse  shaft,  construction  of,  100. 

Reverse-shaft  counterbalance,  11(7. 
Reverse  shaft,  diameter  of,  1(17. 

Reverse  shaft,  location  of,  107. 


Reverse-shaft  pin,  100,  106. 

Reverse  shaft,  position  for  full  gear  backward,  106. 

Reverse  shaft,  position  for  full  gear  forward,  106. 

Reverse  shaft,  position  for  midgear,  106. 

Reverse  shaft,  to  find  position  of,  124. 

Revolving  parts  counterbalanced,  243. 

Rhode  Island  Locomotive  Works,  compound  locomotives 

built  by  the,  625. 

Richardson's  balanced  slide-valves,  67. 
Richardson's,  George  W.,  pop  safety-valves,  384. 
Rigid  wheel  base,  5. 

Rim  of  wheel,  area  of  cross-section  of,  256. 
Rim  of  wheel,  center  of  gravity  of  lead  in,  257. 
Rim  of  wheel,  effect  of  lead  counterbalance  in,  257. 
Rim  of  wheel,  lead  counterbalance  in,  255. 
Ring  for  dome-top,  381. 
Ring  for  fire-box,  434. 
Ring  in  smoke-boxes,  475. 
Ring  on  front  flue-sheet  for  dry  pipe,  347. 
Rings  in  stuffing-boxes,  179. 
Riveted  joints,  446. 
Riveted  joints,  comparison  of  the  strength  of  solid  plate 

with  that  of,  457. 

Riveted  joints,  effects  of  tension  on,  448. 
Riveted  joints,  efficiency  of  single,  458. 
Riveted  joints,  friction  in,  451. 
Riveted  joints,  stress  in,  451. 
Riveted  joints  with  welt  pieces,  459. 
Riveted  lap-joints,  double,  454. 
Riveted  lap-joints,  single,  447. 

Rivets  and  studs  in  dry  and  T-pipe  connections,  349. 
Rivets,  diameter  of,  450. 
Rivets  from  edge  of  plate,  distance  of,  448. 
Rivets  in  fire-box,  435. 

Rivets,  pitch  of,  in  double-riveted  lap-joints,  455. 
Rivets,  pitch  of,  in  single-riveted  lap-joints,  450. 
Rivets,  stress  in  boiler,  452. 
Rivets  through  dry  pipe  and  sleeves,  346. 
Rocker-arms,  computation  of  thickness  of,  77. 
Rocker-arms,  computation  of  width  of,  77. 
Rocker-arms,  hub  on,  78. 
Rocker-arms,  length  of,  76. 
Rocker-arms,  taper  of  holes  in,  74. 
Rocker,  location  of,  118. 
Rocker,  object  of,  36. 
Rocker,  off-set  in,  117,  119. 
Rocker-pins,  74. 
Rockers,  74. 

Rockers,  dimensions  of,  78. 
Rocker-shaft,  computation  of  diameter  of,  75. 
Rocker-shaft,  diameter  of,  74. 
Rocker-shaft,  length  of,  74. 
Rockers,  stress  in,  74. 
Rocker,  the  force  it  must  overcome,  75. 
Rocking  grate-bars  for  burning  soft  coal,  484. 
Rod  brasses,  315. 
Rod  brasses,  babbitted,  315. 
Rod  brasses,  flanges  on,  316. 
Rod  brasses,  metal  in,  317. 
Rod  for  feed-cock,  394. 
Rods  for  piston,  manner  of  fastening,  145. 
Rods  for  pistons,  stress  in,  144. 
Roller  between  spring-saddle  and  spring,  203. 
Roller  valves,  65. 
Rolling  and  axle  friction,  5. 

RULES. 

Rule  1.  To  compute  the  force  an  engine  must  exert  to 
haul  a  train,  7. 

2.  To  compute  number  of  driving  wheels,  8. 

3.  To  compute  the  tractive  force,  17. 

4.  To  compute  thickness  of  cylinder  walls.  'J.">. 

5.  To  compute  steam-port  urea,  27. 

6.  To  compute  steam-port  area  with  aid  of  table.  -_'S. 

7.  To  compute  steam-pipe  area,  .'(0. 

8.  To  compute  stc»m-pipc  area  with  aid  of  table,  31. 

9.  To  compute  piston  speed,  33. 


650 


INDEX. 


Rule  10.  To  compute  travel  of  valve  without  lap,  40. 

"     11.  To  compute  travel  of  valve  without  lap,  40. 

"     12.  To  compute  travel  of  valve  with  lap,  41. 

"     13.  To  compute  travel  of  valve  with  lap,  41. 

"     14.  To  compute  diameter  of  rocker-shaft,  76. 

"     15.  To  compute  width  of  rocker-arms,  77. 

"     16.  To  compute  thickness  of  rocker-arms,  77. 

"     17.  To  find  center  line  of  motion  of  valve  gear,  112. 

"     18.  To  compute  diameter  of  piston-rods,  147. 

"     19.  To  compute  thrust  of  connecting-rod,  160. 

"     20.  To  compute  thrust  of  a  locomotive  connecting- 
rod,  161. 

"     21.  To  compute  approximately  thrust  against  guides, 
161. 

"     22.  To  compute  thrust  against  guide  for  any  position 
of  connecting-rod,  162. 

"    23.  To  compute  area  of  sliding  surface  of  crosshead 
gibs,  163. 

"     24.  To  compute  approximately  area  of  sliding  sur- 
face of  crosshead  gibs,  164. 

"     25.  To  compute  length  and  breadth  of  crosshead 
gibs,  164. 

"     26.  To  compute  diameter  of  crosshead  hubs,  165. 

"     27.  To  compute  thickness  of  guides,  166. 

"     28.  To  compute  depth  of  single  guides,  168. 

"     29.  To  compute  diameter  and  length  of  crosshead- 
pin,  174. 

"    30.  To  compute  diameter  and  length  of  crosshead- 
pin,  175. 

"    31.  To  compute  thickness  of  walls  for  stuffing-box. 
176. 

"    32.  To  compute  thickness  of  flange  of  stuffing-box, 
177. 

"    33.  To  compute  thickness  of  packing  in  stuffing- 
box,  178. 

"    33a.To  compute  cross-sectional  area  of  upper  engine 
frame-braces,  192. 

"     34.  To  compute  thickness  of  flange  on  gland,  179. 

"     34a.To  compute  depth  of  upper  engine  frame-brace, 
193. 

"    35.  To  compute  diameter  of  studs  for  stuffing-boxes, 
180. 

"    35o.To  compute  depth  of  lower  engine  frame-brace, 
194. 

"    36.  To  compute  depth  of  lower  engine  frame-brace, 
194. 

"    37.  To  compute  depth  of  frame-brace,  slab  form, 
197. 

"    38.  To  compute  diameter  of  bolts  through  frame 
and  splice,  198. 

"    39.  To  compute  number  of  bolts  through  frame  and 
splice,  198. 

"    40.  To  compute  depth  of  front  frame-splice,  199. 

"    41.  To  compute  diameter  of  driving-axle  journal, 
209. 

"    42.  To  compute  weight  of  a  simple  counterbalance, 
225. 

"    43.  To  compute  weight  of  crank  referred  to  crank- 
pin,  229. 

"    44.  To  compute  weight  of  counterbalance,  235. 

"    45.  To  compute  number  of  cubic  inches  in  counter- 
balance, 235. 

"    46.  To  compute  thickness  of  counterbalance,  235. 

"    47.  To  find  common  center  of  gravity  of  three  seg- 
ments in  counterbalance,  238. 

"    48.  To  find  common  center  of  gravity  of  four  seg- 
ments in  counterbalance,  238. 

"    49.  To  find  common  center  of  gravity  of  five  seg- 
ments in  counterbalance,  239. 

"    50.  To  find  common  center  of  gravity  of  any  two 
bodies,  240. 

"     51.  To  compute  area  of  irregular  surfaces,  254. 

"     52.  To  compute  weight  of  lead  in  rim  of  wheels, 
255. 

"    53.  To  compute  center  of  gravity  of  lead  in  rim  of 
wheel,  257. 

"    54.  To  compute   cross-sectional   area  of  main-rod 
bolts,  276. 


Rule 

55. 

tt 

56. 

it 
ti 
it 

57. 
58. 
59. 

tt 

60. 

11 

61. 

tt 

62. 

It 

63. 

tt 

64. 

It 

65. 

tt 

66. 

tt 

67. 

tt 

68. 

ft 

69. 

tt 

70. 

tt 

71. 

tt 

72. 

ft 

73. 

tt 

74. 

It 

75. 

tt 

76. 

tt 

77. 

n 

78. 

tt 

79. 

tt 

80. 

(t 

81. 

It 

82. 

It 

83. 

ft 

84. 

tl 

85. 

tt 
tt 
tt 
tt 
ft 
tl 

86. 
87. 
88. 
89. 
90. 
91. 

tl 

92. 

it 

93. 

To  compute  diameter  of  bolts  through  rear 
strap  of  main-rod,  276. 

To  compute  number  of  bolts  through  rear  strap 
of  main-rod,  276. 

To  compute  thickness  of  main-rod  straps,  285. 

To  compute  thickness  of  side-rod  straps,  288. 

To  compute  thickness  of  side-rod  straps,  Mogul 
engines,  292. 

To  compute  thickness  of  side-rod  strap,  main- 
pin,  Mogul  engines,  292. 

To  compute  thickness  of  front  and  rear  side-rod 
straps,  consolidation  engine,  295. 

To  compute  thickness  of  central  side-rod  straps, 
consolidation  engine,  295. 

To  compute  diameter  of  side-rod  bolts  for  eight- 
wheeled  engines,  298. 

To  compute  diameter  of  ride-rod  bolts  for  Mogul 
engines,  298. 

To  compute  diameter  of  side-rod  bolts  for  con- 
solidation engines,  299. 

To  compute  area  of  smallest  transverse  section 
of  main-rod,  301. 

To  compute  thickness  and  depth  of  main-rod, 
302. 

To  compute  cross-sectional  area  of  side-rods  for 
eight-wheeled  engines,  308. 

To  compute  thickness  and  depth  of  side-rods 
for  eight-wheeled  engines,  309. 

To  compute  depth  for  front  side-rod,  Mogul  en- 
gine, 310. 

To  compute  cross-sectional  area  of  central  side- 
rod,  consolidation  engine,  312. 

To  compute  thickness  and  depth  of  central  side- 
rod,  consolidation  engine,  312. 

To  compute  thickness  and  depth  of  rear  and 
front  side-rods  for  consolidation  engines,  312. 

To  compute  projected  area  of  main-rod  crank- 
pin  journal  for  eight-wheeled  engines,  320. 

To  compute  diameter  and  length  of  main-rod 
crank-pin  journal  for  eight-wheeled  engines, 
320. 

To  compute  projected  area  of  main  crank-pin 
side-rod  journal  for  eight-wheeled  engines, 
320. 

To  compute  diameter  and  length  of  main  crank- 
pin  side-rod  journal  for  eight-wheeled  en- 
gines, 321. 

To  compute  diameter  and  length  of  main  crank- 
pin  journal  for  Mogul,  ten-wheeled,  and  con- 
solidation engines,  329. 

To  compute  diameter  and  length  of  side-rod 
journals  on  main  crank-pin  for  Mogul,  ten- 
wheeled,  and  consolidation  engines,  330. 

To  compute  projected  area  of  front  and  rear 
side-rod  pins  in  Mogul  and  ten-wheeled  en- 
gines, 333. 

To  compute  diameter  and  length  of  front  and 
rear  side-rod  pins  in  Mogul  and  ten- wheeled 
engines,  333. 

To  compute  projected  area  of  side-rod  pins  for 
consolidation  engines,  336. 

To  compute  diameter  and  length  of  side-rod 
pins  for  consolidation  engines,  336. 

To  compute  projected  area  of  knuckle-joint  pin 
in  side-rods,  340. 

To  compute  diameter  and  length  of  knuckle- 
joint  pin  in  side-rods,  341. 

To  compute  lift  of  throttle- valve,  358. 

To  compute  tension  on  safety-valve  springs,  375. 

To  compute  length  of  safety-valve  lever,  376. 

To  compute  steam  pressure  on  safety-valve,  376. 

To  compute  area  of  safety-valve,  376. 

To  compute  distance  from  center  of  safety-valve 
to  fulcrum,  377. 

To  compute  steam  pressure  on  safety-valve, 
weight  of  lever  considered,  379. 

To  compute  area  of  safety-valve,  weight  of 
lever  considered,  379. 


651 


Rule  94.  To  compute  distance  from  fulcrum  to  center  of 
valve,  Wright  of  lever  considered,  :J79. 

••     !i.V   To    compute    tension    on    safety-valve    spring, 
weight  of  lever  considered,  380. 

"     96.  To  compute  cross-sectional  area  of  full-stroke 
pump  plungers,  397. 

"    97.  To  compute  cross-sectional  area  of  short-stroke 
pump  plungers,  398. 

"     98.   To  compute  depth  of  equalizing  lever  for  driv- 
ing-wheel springs,  406. 

"    99.  To  compute  cross-sectional  area  of  equalizing- 
lever  fulcrum,  409. 

"     100.  To   compute    cross-sectional    area   of  spring 
hangers,  410. 

"    101.  To  compute  number  of  leaves  in  elliptical 
springs,  415. 

"     102.  To  compute  deflection  of  elliptical  springs,  416. 

"     103.  To  compute  area  of  grate  surface,  soft -coal, 
422.  .     • 

"     104.  To  compute  area  of  grate  surface,  hard-coal, 
41'ii. 

"     105.  To  compute  space  between  boiler  tubes,  443. 

"     106.  To  compute  stress  in  stay-bolts,  480. 

"     107.  To  compute  distance  between  centers  of  stay- 
bolts,  480. 

"     108.  To  compute  thickness  of  stayed  sheets,  481. 

"     109.  To  compute  depth  of  grate-bars  for  burning 
wood,  484. 

"     110.  To  compute  area  of  orifices  in  double-exhaust 
nozzles,  507. 

"     111.  To  compute  area  of  orifices  in  single-exhaust 
nozzles,  508. 

"     112.  To  compute  depth  of  bumper  beam,  522. 

"     113.  To  compute  length  of  engine  pony-truck,  544. 

"     114.  To  compute  diameter  and  length  of  engine- 
truck  axle-journals,  550. 

"     115.  To  compute  diameter  of  cylinders  for  simple 
engines,  595. 

"     116.  To  compute  required  mean  effective  pressure, 
596. 

"    117.  To  compute  ratio  of  expansion,  clearance  neg- 
lected, 600. 

"     118.  To  compute  terminal  pressure,  601. 

"     119.  To  compute  mean  pressure,  without  clearance, 
602. 

"     120.  To  compute  ratio  of  expansion,  with  clearance, 
604. 

"    121.  To  compute  mean  pressure,  with  clearance, 
605. 

"     122.  To  compute  horse-power,  605. 

"     123.  To  compute  normal  pressure,  611. 

"     124.  To  compute  force  required  to  haul  a  train  on 
a  grade,  613. 

"     125.  To  compute  force  required  to  haul  a  train  on 
a  grade,  (ill!. 

"     126.  To  compute  resistance  due  to  speed,  614. 

"     127.  To  compute  resistance  due  to  curves,  615. 

"     128.  To  compute  point  of  exhaust  closure,  619. 

"     129.  To  compute  steam-port  area  in  high-pressure 

cylinders  in  compound  engines,  627. 
"     130.  To  compute  steam-port  area  in  low-pressure 

cylinders  in  compounds,  0:27. 

"     131.  To  compute  diameter  of  cylinders  for  four- 
cylinder  compounds,  632. 

Rules,  formulas,  and  data,  595. 
Running-board  brackets,  531. 
Running  board,  position  of,  532. 

8 

Saddle  for  link,  87,  100. 
Saddle-pin,  87,  100. 
Saddle-pin  for  link,  length  of,  106. 
Saddle  pin,  to  find  position  of,  123. 

Saddles  for  cylinders,  20. 

Saddle,  thickness  of  metal  in  sides  of,  25. 

Safety  chains  for  tender  trucks,  564. 


Safety  links  between  engine  and  tender,  528. 

Safety  valve,  area  of,  376,  379. 

Safety-valve  attachments,  371. 

Safety-valve  bearing  surface,  angle  of,  371. 

Safety  valve,  Crosby,  385. 

Safety  valve,  distance  from  fulcrum  to  center  of,  377, 

380. 

Safety  valve,  handle  for  opening,  362. 
Safety-valve  lever,  373. 
Safety-valve  lever,  length  of,  374,  376,  380. 
Safety-valve  levers,  center  of  gravity  of,  377. 
Safety-valve  opening  in  top  of  dome,  diameter  of,  373. 
Safety  valve,  pop,  373,  380,  384. 
Safety  valves,  common,  371. 
Safety  valves,  computations  relating  to,  374. 
Safety  valves,  diameter  of  pop,  384. 
Safety  valves,  encased  pop,  385. 
Safety  valves,  number  of,  362. 
Safety  valves,  position  of,  380. 
Safety  valves,  pressure  on  springs  of  pop,  386. 
Safety-valve  spring  balance,  362,  369,  370. 
Safety-valve  springs,  tension  on,  374,  375,  380. 
Safety  valves,  steam  pressure  on,  376,  379. 
Sand-boxes,  510. 
Sand-boxes,  construction  of,  511. 
Sand,  condition  and  use  of,  511. 
Sand-pipes,  511. 
Sand-pipes,  position  of,  513. 
Schenectady  Locomotive  Works,  compound  engines  built 

by  the,  620. 

Scoops,  tenders  with  water,  572. 
Segment,  form  of  templet  for,  246. 
Segments  in  counterbalance,  number  of,  244. 
Segments  in  counterbalance,  use  of  two,  235. 
Segments  of  a  counterbalance,  to  find  the  common  cen- 
ter of  gravity  of  two,  236. 
Segments  of  counterbalance,  bolts  for,  242. 
Segments  of  counterbalance,  common  center  of  gravity 

of  five,  239. 
Segments  of  counterbalance,  common  center  of  gravity 

of  four,  238. 
Segments  of  counterbalance,  common  center  of  gravity 

of  three,  238. 

Segments  of  counterbalance,  form  of,  242. 
Segments,  two  in  a  counterbalance,  thickness  of,  237. 
Segments,  two  in  a  counterbalance,  to  find  weight  of 

each,  236. 

Selection  of  iron  for  cylinders,  23. 
Separator  in  throttle  pipes,  351. 
Set-screws  in  eccentric,  81. 

Setting  a  simple  valve  gear,  practical  method  of,  96. 
Setting  of  eccentrics,  practical  way,  47. 
Setting  the  valve  gear  of  a  locomotive,  practical  example, 

125. 

Shaft  for  drop-plate  in  furnace,  486. 
Shaft,  lifting,  object  of,  36. 
Shaking  lever  for  grate  bars,  486. 
Shaking-lever  handle,  486. 
Sheets  for  bumpers,  522. 
Sheets,  thickness  of  stayed,  481. 
Shifting  engine,  six-wheeled,  built  by  the  Pennsylvania 

R.  R.,  589. 

Shifting  link,  change  of  lead  with,  93. 
Shifting  link  for  American  locomotives,  88. 
Shifting  links,  88. 

Shifting  links,  angular  advance  of  eccentric  for,  88. 
Shifting  links,  curvature  of,  88. 
Shifting  links,  lead  variable  with,  88. 
Shimming  pieces  under  tires,  25!i. 
Shoe  for  frame  pedestals,  184. 
Short  double-exhaust  pipes,  504. 
Short  exhaust  pipes,  where  used,  506. 
Short  single-exhaust  pipes,  505. 

Short-stroke  pump  plungers,  cross-sectional  area  of,  398.r' 
Short-stroke  pumps,  .'J87,  395. 
Short  wedge  for  frame  pedestals,  1  sf>. 
Shrinkage  allowance  for  driving-wheel  tires,  217. 
Shrinking  tires  on  wheels,  215. 


652 


INDEX. 


Side-bearing  tender  trucks,  564. 

Side-rod  bolts  for  consolidation  engines,  diameter  of,  298. 

Side-rod  bolts  for  eight-wheeled  engines,  cross-sectional 

area  of,  297. 
Side-rod  bolts  for  eight-wheeled  engines,   diameter  of, 

297. 

Side-rod  bolts  for  eight-wheeled  engines,  stress  in,  297. 
Side-rod  bolts  for  Mogul  engines,  diameter  of,  298. 
Side-rod  bolts  for  ten-wheeled  engines,  diameter  of,  298. 
Side-rod  bolts,  strength  of,  296. 
Side-rod  brasses,  315. 
Side-rod  I  section,  solid  ends,  270. 
Side-rod  pin  for  a  consolidation  engine,  328,  335. 
Side-rod  pins  for  eight-wheeled  engines,  320. 
Side-rod  pins  for  Mogul  and  ten-wheeled  engines,  332. 
Side-rod  pins,  wheel  fit,  337. 
Side-rods  and  main-rods,  267. 
Side-rods,  buckling  of,  306. 

Side-rods,  conditions  under  which  they  work,  305. 
Side-rods,  cross-sectional  area  of,  308. 
Side-rods  for  consolidation  engines,  289,  312. 
Side-rods  for  eight-wheeled  engines,  position  of,  290. 
Side-rods  for  eight-wheeled  engines,  thickness  and  depth 

of,  309. 

Side-rods  for  engines  on  elevated  roads,  278. 
Side-rods,  form  of  keys  for,  299,  300. 
Side-rods  for  Mogul  and  ten-wheeled  engines,  309. 
Side-rods  for  Mogul  engines,  281. 

Side-rods  for  narrow-gauge  eight-wheeled  engines,  290. 
Side-rods,  four-wheeled  connected,  keys  for,  270. 
Side-rods,  knuckle  joint  for,  338. 
Side-rods,  length  of,  how  taken,  317. 
Side-rods,  liners  for,  270. 

Side-rods,  number  and  diameter  of  bolts  in,  267. 
Side-rods,  pin  through  knuckle  joint  of,  281. 
Side-rods,  proportions  of,  308. 
Side-rods,  ratio  of  thickness  and  depth  of,  309. 
Side-rods,  types  of,  307. 
Side  rods  with  an  I  cross-section,  307. 
Side-rod  with  solid  end,  327. 
Side  sheets,  inclination  of  furnace,  432. 
Sight -feed  lubricators,  556. 
Simple  valve  gear,  36. 
Single-exhaust  pipes,  advantages  of,  506. 
Single-riveted  lap  joints,  447. 
Size  of  cylinders,  11. 
Size  of  eccentric-rod  pins,  103. 
Size  of  pumps,  396. 
Skeleton  link,  99. 
Skeleton  links,  difficulty  with,  103. 
Slab  frame  brace,  cross-sectional  area  of,  197. 
Slab  frame  brace,  depth  of,  197. 
Slab  frame  braces,  least  thickness  of,  197. 
Slab  frames,  196. 
Sleeve  on  dry  pipes,  346. 
Slide  throttle  valve,  343. 
Slide-valve,  Allen,  68. 

Slide-valve,  Allen,  advantages  claimed  for,  69. 
Slide-valve,  clearance  of,  59. 
Slide-valve,  conditions  it  must  fulfill,  38. 
Slide-valve,  friction  of,  64. 
Slide-valve,  inside  lap  of,  59. 
Slide-valve,  inside  lead  of,  59. 
Slide-valve,  power  required  to  work  a,  63. 
Slide-valve,  practical  construction  of,  57. 
Slide-valve,  pressure  on,  64,  65. 
Slide-valve,  primitive,  48. 
Slide-valve,  problems  relating  to  lap  of,  52. 
Slide-valve,  purpose  of  giving  lap  to,  49. 
Slide-valves,  34. 
Slide-valves,  balanced,  66. 
Slide-valves,  balanced,  Richardson's,  67. 
Slide-valves,  classification  of,  39. 
Slide-valves,  duty  of,  35. 
Slide-valves,  hole  in  top  of  balanced,  67. 
Slide-valves,  problems  relating  to,  reduced  to  simplest 

form,  35. 
Slide-valves,  roller,  65. 


Slide-valves,  thickness  of  metal  in,  34,  59. 

Slide-valve,  to  find  lap  and  travel  of,  55,  56. 

Slide-valve,  travel  of,  37,  82. 

Slide-valve,  travel  of,  affects  point  of  cut-off,  54. 

Sliding  surface  of  crosshead  gibs,  area  of,  163. 

Slip  of  link-blocks,  102. 

Sloping  crown  sheet,  430. 

Smoke-box  doors,  476. 

Smoke-boxes,  braces  from  frames  to,  522. 

Smoke-boxes,  capacity  of,  473. 

Smoke-boxes,  long,  blowing  out  by  steam,  474. 

Smoke-boxes,  long,  cast-iron  caps  on  side  of,  474. 

Smoke-boxes,  long,  cinder-box  on,  474. 

Smoke-boxes,  long,  construction  of,  473. 

Smoke-boxes,  long,  diaphragm  plates  in,  474. 

Smoke-boxes,  long,  or  extension  fronts,  472. 

Smoke-boxes,  long,  position  of  diaphragm  plates  in,  474. 

Smoke-boxes,  New  York  Elevated  Railroad,  471. 

Smoke-boxes,  short,  470. 

Smoke-boxes,  short,  petticoat  pipe  in,  470. 

Smoke-box  fronts,  476. 

Smoke-box  rings,  475. 

Smoke-box  shell,  construction  of,  475. 

Smoke-box  shell,  length  of,  475. 

Smoke-box  shell,  number  of  courses  in,  475. 

Soft-coal  burning  engines,  position  of  fire-box  for,  188. 

Soft-coal  burning  engines,  space  required  between  axles 
in,  188. 

Solid  link,  99. 

Solid  pistons,  140. 

Space  between  tubes,  443. 

Space  for  eccentrics,  82. 

Space  for  steam  and  exhaust  passages  is  limited,  33. 

Specification  for  Pennsylvania  R.  R.,  eight-wheeled  en- 
gine, 585. 

Specification  for  Pennsylvania  R.  R.,  shifting  engine,  589. 

Speed  of  piston,  to  determine,  33. 

Speed,  resistance  due  to,  614. 

Spider  for  pistons,  138. 

Spider  for  pistons,  depth  of,  139. 

Spiral  punch,  injury  to  plates,  451. 

Splice  and  frame,  diameter  of  bolts  through,  198. 

Splice  and  frame  forged  in  one  piece,  201. 

Splice,  casting  for  front  end  of  frame,  200. 

Splice,  depth  of  front  frame,  199. 

Splice  for  consolidation  and  Mogul  engine  frames,  200, 
201. 

Splice  for  frames,  form  of  front  end  of,  200. 

Splices  for  frames,  passenger  engines,  187,  197. 

Splices  for  frames,  recess  for  cylinder  saddle  in,  199. 

Spring  balance  for  safety  valve,  362,  369,  370. 

Spring  equalizing  lever,  load  on  driving  wheel,  403. 

Spring  equalizing  levers,  purpose  of,  400. 

Spring  gear,  depth  of  equalizing  lever  for,  402,  405. 

Spring  gear,  equalizing  lever  for  driving  wheel,  400. 

Spring  gear  for  driving  wheels,  eight- wheeled  engine,  400. 

Spring  gear  for  driving  wheels,  ten-wheeled  engines,  400. 

Spring  gear  for  narrow-gauge  eight-wheeled  engine,  403. 

Spring  gear,  fulcrum  for  driving  wheel,  400,  409. 

Spring  gear,  load  on  fulcrum  for  driving  wheel,  403. 

Spring  gear,  Pennsylvania  R.  R.,  414. 

Spring  gears  for  consolidation  engines,  driving  wheel, 
408. 

Spring  gears  for  Mogul  engines,  driving  wheel,  408. 

Spring  gear,  stress  in  equalizing  fulcrum,  409. 

Spring  gear,  thickness  of  equalizing  lever  for,  402,  408. 

Spring  hangers  for  driving-wheel  springs,  400,  410. 

Spring  hangers,  ratio  of  thickness  to  width,  411. 

Spring  hangers,  stress  per  square  inch,  410. 

Spring  hangers,  various  forms  of,  411. 

Spring  hanger,  tension  on  driving  wheel,  403. 

Springing  of  the  valve-rod,  76. 

Spring  saddles  for  driving-wheel  springs,  202,  400. 

Springs  for  balanced  slide-valves,  66. 

Springs  for  driving  wheels,  414. 

Springs  for  driving  wheels,  deflection  of,  416. 

Springs  for  driving  wheels,  length  of,  414. 

Springs  for  driving  wheels,  number  of  blades  in,  415. 


I\IIK.\. 


653 


Springs  for  driving  wheels,  set  of,  415. 

Springs  for  driving  wheels,  thickness  of  blades  of,  415. 

Springs  for  lifting-slmft  countcrlmlHiice,  107. 

Springs  for  lifting-shaft   counterbalance,  dimensions  of, 

ins. 
Springs,  front,  in  ten-wheeled,  Mogul,  and  consolidation 

engines,  load  on,  546. 
Springs,  helical,  compression  of,  386. 
Springs,  helical,  si/.e  of  steel  for,  386. 
Springs,  loud  on  driving  wheel,  402. 
Stack,  diamond,  with  donlile  shell,  496. 
Stack,  diamond,  with  single  shell,  497. 
Stack,  H.  A.  Luttgens.  :>ni. 
Stack,  Radley  &  Hunter,  498. 
Stacks,  diameter  of,  502. 
Stacks,  form  of  saddles  for,  502. 
Stacks,  length  of,  503. 
Stack,  straight,  499. 
Stack,  tapered.  499. 
Stack,  wood-burning,  497. 
Standard  diameters  of  wheel  centers,  217. 
Standard  tender-truck  axle-boxes,  579. 
Standard  tender-truck  axles,  581. 
Stand  for  bells,  513. 
Stand  pipe  for  throttle  valve,  344. 
Stand  to  receive  boiler  cocks,  valves,  etc.,  366. 
Starting  valve  for  four-cylinder  locomotives,  8.  M.  Vau- 

clain,  631. 
Stationary  links,  88. 

Stationary  links,  angular  advance  of  eccentric  for,  88,  90. 
Stationary  links,  curvature  of,  88. 
Stationary  links,  lead  constant  with,  88,  89. 
Stationary  links,  length  of  radius  for,  90. 
Stationary  links,  to  find  angular  advance  for,  91. 
Stay  bolts,  distance  between  centers  of,  480,  481. 
Stay  bolts  for  crown  sheets,  radial,  432. 
Stay  bolts  in  Belpaire  boilers,  464. 
Stay  bolts  in  boilers,  463. 
Stay  bolts  in  boilers,  hollow,  463. 
Stay  bolts,  radial,  to  crown  sheet,  463. 
Staying  back  head  of  boiler,  469. 
Staying  front  tube  sheet,  469. 
Steam  accounted  for  by  the  indicator,  607. 
Steam,  back  pressure,  602. 

Steam  chamber  for  receiving  valves  on  boiler,  561. 
Steam-chest  seats,  21. 
Steam-chests,  vacuum  valves  in,  67. 
Steam,  compression  of,  in  compounds,  618. 
Steam,  diagram  showing  events  of  distribution  of,  63. 
Steam,  events  of  distribution  of,  60. 
Steam,  expansive  working  of,  in  compounds,  619. 
Steam-gauge  lamp  bracket,  362. 
Steam-gauge  stand,  362,  369. 
Steam-gauge   stand  and  chamber  arranged   to  receive 

valves,  cocks,  etc.,  366. 
Steam-gauge  stand,  position  of,  373. 
Steam,  lap  required  for  expansion  of,  49. 
Steam  line,  607. 
Steam  passage,  21. 
Steam  passage    divided    into  two  branches,  advantage 

gained,  22. 

Steam  passage,  duty  of,  22. 
Steam  passage  openings,  size  of,  32. 
Steam  passages,  small  space  for,  33. 
Steam  passage,  thickness  of  metal  around,  25. 
Steam-pipe  area,  31. 

Steam-pipe  areas,  table  of  proportional,  31. 
Steam  pipes,  30. 
Steam  pipes,  ball  joint  for,  .'13. 
Steam  pipes,  thickness  of,  33. 
Steam  pipes,  to  compute  area  of,  30. 
Steam  pipes,  to  compute  area  of,  with  aid  of  table,  31. 
Steam-port  area  in  compounds,  627. 
Steam-port  areas,  table  of.  L'H. 
Steam-port  area,  tn  compute.  27. 
Steam  ports,  22. 
Steam  ports,  area  .>)',  26. 
Steam  ports,  length  and  breadth  of,  26. 


Steam  pressure,  absolute,  597. 

Steam  pressure,  computation  at  any  point  of  stroke, 

598. 
Steam  pressure,  computation  of  mean  effective,  with  aid 

of  ordinates,  607. 
Steam  pressure,  computation  of  mean  effective,  with  aid 

of  planimeter,  607. 
Steam  pressure,  computation  of  required  mean  effective, 

596. 

Steam  pressure,  computation  of  terminal,  599,  601. 
Steam  pressure  for  a  given  cut-off,  mean  effective,  597, 

602,  605. 

Steam  pressure,  initial,  597. 
Steam  pressure,  mean,  601. 
Steam  pressure,  mean,  computation  of,  602. 
Steam  pressure,  terminal,  598. 
Steam  pressure  which  boilers  will  stand,  formulas  for, 

462. 

Steam  space  in  boiler,  443. 

Steam,  temperature  falls  with  the  expansion  of,  617. 
Steam,  to  find  point  of  admission,  60. 
Steam,  to  find  point  of  compression  of,  60. 
Steam,  to  find  point  of  cutting  off,  60. 
Steam,  to  find  point  of  release,  60. 
Steam  ways,  duty  of,  22. 
Steel  crank-pins,  317. 
Steel  fire-boxes,  462. 
Steel  links,  102. 
Stephenson's  link  motion,  86. 
Stephenson's  link,  object  of,  36. 
Straight  legs  in  engine  pedestal,  position  of,  190. 
Straps  for  eccentric,  form  of,  81. 
Straps  for  eccentrics,  79. 
Straps  for  eccentrics,  joints  in,  81. 
Straps  for  main-  and  side-rods,  proportions  of,  270. 
Straps  for  main-rods,  diameter  of  bolts  through  front, 

279. 

Straps  for  main-rods,  diameter  of  bolts  through  rear,  276. 
Straps  for  main-rods,  number  of  bolts  through  rear,  276. 
Straps  for  side-rods,  consolidation  engines,  294. 
Straps  for  side-rods,    eight-wheeled  passenger  engine, 

thickness  of,  286. 
Straps  for  side-rods  for  consolidation  engines,  thickness 

of,  295. 

Straps  for  side-rods  for  Mogul  engines,  290. 
Straps  for  side-rods  for  Mogul  engines,  thickness  of,  292. 
Straps,  least  thickness  of  side-rod,  294. 
Straps,  oil-holes  in  main-  and  side-rod,  296. 
Straps,  thickness  of  main-rod,  283. 
Strength  of  crank-pins,  318. 
Strength  of  crosshead-pins,  1 74. 
Strength  of  crown  bars,  computation  of,  466. 
Strength  of  iron  plates,  tensile,  451,  462. 
Strength  of  piston  keys,  148. 
Strength  of  plates,  apparent,  451. 
Strength  of  plates  restored  by  annealing,  451. 
Strength  of  plates  restored  by  reaming  the  holes,  451. 
Strength  of  riveted  joints  compared  with  that  of  solid 

plate,  457. 

Strength  of  steel  plates,  tensile,  462. 
Stress  in  boiler  rivets,  452. 
Stress  in  boiler  shell,  longitudinally,  447. 
Stress  in  boiler  shell,  transverse,  447. 
Stress  in  bolts  through  frame  and  splice,  198. 
Stress  in  bumper  beams,  520. 
Stress  in  crosshead  hubs,  165. 
Stress  in  cylinder-head  bolts,  23. 
Stress  in  oblique  braces,  481. 
Stress  in  piston-rods,  144. 
Stress  iii  riveted  joints,  451. 
Stress  in  rockers,  74. 
Stress  in  rod-bolts,  shearing,  272. 
Stress  in  side-rod  straps,  eight-wheeled  engines,  288. 
Strc>s  in  side-rod  straps  for  Mogul  engines,  291. 
Stress  cm  slay  holts,  479. 

Stress  per  square  inch  in  engine  frame  braces,   l!(l. 
Stress  per  square  inch  of  cross-sectional  area  of  main- 
rod  straps,  2x."i. 


654 


INDEX. 


Stress  per  square  inch  of  section  of  rod-bolts,  allowable, 

275. 

Stroke,  backward,  89. 
Stroke,  forward,  89. 
Stroke  of  pump,  to  find,  398. 

Studs  and  rivets  in  dry  and  T-pipe  connections,  349. 
Studs  for  glands  and  stuffing-boxes,  180. 
Studs  for  throttle-valve  stuffing-box,  353. 
Studs  in  cylinder  head,  24. 
Stuffing-box,  brass  ring  in,  179. 
Stuffing-boxes  and  glands,  proportions  of,  176. 
Stuffing-box  for  throttle-valve  rod,  352,  355. 
Stuffing-box  for  throttle-valve  rod,  position  of,  355. 
Stuffing-box  gland  for  throttle-valve  rod,  355. 
Suction  hose,  393. 
Support  for  pulling-bar,  528. 
Supports  for  boilers,  522. 
Supports  of  engines,  points  of,  408. 
Surface  of  counterbalance,  area  of,  249. 
Surfaces,  area  of  irregular,  254. 
Suspension  of  link,  100. 
Symmetrical  motion,    eccentric-rods   of   infinite   length 

give,  37. 
Symmetrical  motion  of  piston,  how  to  obtain,  50. 


TABLES. 

Table  1.  Diameter  of  driving  wheels,  eight-wheeled  loco- 
motive, 10. 

"  2.  Diameter  of  driving  wheels,  Mogul  locomotive, 
10. 

"  3.  Diameter  of  driving  wheels,  ten-wheeled  loco- 
motive, 10. 

"  4.  Diameter  of  driving  wheels,  consolidation  loco- 
motive, 10. 

"  5.  Weight  and  hauling  capacity  of  eight- wheeled 
locomotives,  17. 

"  6.  Weight  and  hauling  capacity  of  Mogul  locomo- 
tives, 18. 

"  7.  Weight  and  hauling  capacity  of  ten- wheeled 
locomotives,  18. 

"  8.  Weight  and  hauling  capacity  of  consolidation 
locomotives,  19. 

"      9.  Proportional  steam-port  area,  28. 

"     10.  Proportional  steam-pipe  areas,  31. 

"     11.  Size  of  steam  and  exhaust  openings,  32. 

"     12.  Proportional  dimensions  of  eccentrics,  85. 

"     13.  Breadth  and  thickness  of  link,  105. 

"     14.  Diameter  of  piston-rods,  148. 

"     15.  Average  dimensions  of  crosshead-pins,  173. 

"     16.  Computed  dimensions  of  crosshead-pins,  175. 

"  17.  Dimensions  of  driving-axle  journals  for  Mogul 
engines,  211. 

"  18.  Dimensions  of  driving-axle  journals  for  ten- 
wheeled  engines,  211. 

"  19.  Dimensions  of  driving-axle  journals  for  con- 
solidation engines,  212. 

"  20.  Dimensions  of  driving-axle  journals  in  actual 
service,  212. 

"     21.  Standard  sizes  of  wheel  centers,  217. 

"  22.  Thickness  and  depth  of  main-rods,  pressure 
120  pounds,  304. 

"  23.  Thickness  and  depth  of  main-rods,  pressure 
130  pounds,  304. 

"  24.  Thickness  and  depth  of  main-rods,  pressure 
140  pounds,  304. 

"  25.  Thickness  and  depth  of  main-rods,  pressure 
150  pounds,  305. 

"  26.  Thickness  and  depth  of  main-rods,  pressure 
160  pounds,  305. 

"  27.  Thickness  and  depth  of  side-rods  for  eight- 
wheeled  engines,  pressure  120  pounds,  310. 

"  28.  Thickness  and  depth  of  side-rods  for  eight- 
wheeled  engines,  pressure  130  pounds,  310. 

"  29.  Thickness  and  depth  of  side-rods  for  eiglit- 
wheeled  engines,  pressure  140  pounds,  311. 


Table  30.  Thickness  and  depth  of  side-rods  for  eight- 
wheeled  engines,  pressure  150  pounds,  311. 

"  31.  Thickness  and  deptli  of  side-rods  for  eight- 
wheeled  engines,  pressure  160  pounds,  311. 

"  32.  Thickness  and  depth  of  side-rods  for  consolida- 
tion engines,  pressure  1UO  pounds,  313. 

"  33.  Thickness  and  depth  of  side-rods  for  consolida- 
tion engines,  pressure  130  pounds,  313. 

"  34.  Thickness  and  deptli  of  side-rods  for  consolida- 
tion engines,  pressure  140  pounds,  314. 

"  35.  Thickness  and  depth  of  side-rods  for  consolida- 
tion engines,  pressure  150  pounds,  314. 

"  36.  Thickness  and  depth  of  side-rods  for  consolida- 
tion engines,  pressure  160  pounds,  314. 

"  37.  Dimensions  of  crank-pin  journals  for  eight- 
wheeled  engines,  pressure  120  pounds,  321. 

"  38.  Dimensions  of  crank-pin  journals  for  eight- 
wheeled  engines,  pressure  130  pounds,  322. 

"  39.  Dimensions  of  crank-pin  journals  for  eight- 
wheeled  engines,  pressure  140  pounds,  322. 

"  40.  Dimensions  of  crank-pin  journals  for  eight- 
wheeled  engines,  pressure  150  pounds,  322. 

"  41.  Dimensions  of  crank-pin  journals  for  eight- 
wheeled  engines,  pressure  160  pounds,  323. 

"  42.  Dimensions  of  main  crank-pin  journals  for  Mo- 
gul, ten-wheeled,  and  consolidation  engines, 
pressure  120  pounds,  331. 

"  43.  Dimensions  of  main  crank-pin  journals  for  Mo- 
gul, ten-wheeled,  and  consolidation  engines, 
pressure  130  pounds,  331. 

"  44.  Dimensions  of  main  crank-pin  journals  for  Mo- 
gul, ten-wheeled,  and  consolidation  engines, 
pressure  140  pounds,  331. 

"  45.  Dimensions  of  main  crank-pin  journals  for  Mo- 
gul, ten-wheeled,  and  consolidation  engines, 
pressure  150  pounds,  332. 

''  46.  Dimensions  of  main  crank-pin  journals  for  Mo- 
gul, ten-wheeled,  and  consolidation  engines, 
pressure  160  pounds,  332. 

"  47.  Dimensions  of  front  and  rear  side-rod  pins  for 
Mogul  and  ten-wheeled  engines,  pressure 
120  pounds,  334. 

"  48.  Dimensions  of  front  and  rear  side-rod  pins  for 
Mogul  and  ten-wheeled  engines,  pressure 
130  pounds,  334. 

"  49.  Dimensions  of  front  and  rear  side-rod  pins  for 
Mogul  and  ten-wheeled  engines,  pressure 
140  pounds,  334. 

"  50.  Dimensions  of  front  and  rear  side-rod  pins  for 
Mogul  and  ten-wheeled  engines,  pressure 
150  pounds,  335. 

"  51.  Dimensions  of  front  and  rear  side-rod  pins  for 
Mogul  and  ten-wheeled  engines,  pressure 
160  pounds,  335. 

"  52.  Dimensions  of  side-rod  pins  for  consolidation 
engines,  pressure  120  pounds,  337. 

"  53.  Dimensions  of  side-rod  pins  for  consolidation 
engines,  pressure  130  pounds,  337. 

"  54.  Dimensions  of  side-rod  pins  for  consolidation 
engines,  pressure  140  pounds,  338. 

"  55.  Dimensions  of  side-rod  pins  for  consolidation 
engines,  pressure  150  pounds,  338. 

"  56.  Dimensions  of  side-rod  pins  for  consolidation 
engines,  pressure  160  pounds,  338. 

"  57.  Dimensions  of  knuckle-joint  pins  for  side-rods, 
pressure  120  to  140  pounds,  341. 

"  58.  Dimensions  of  knuckle-joint  pins  for  side-rods, 
pressure  140  to  160  pounds,  342. 

"  59.  Grate  area  for  soft-coal  burning  eight-wheeled 
engines,  423. 

"  60.  Grate  area  for  soft-coal  burning  consolidation 
engines,  423. 

"  61.  Grate  area  for  hard-coal  burning  eight-wheeled 
engines,  427. 

"  62.  Grate  area  for  hard-coal  burning  consolidation 
engines,  427. 

"     63.  Aggregate  tube  area  as  found  in  practice,  439. 

"     64.  Aggregate  tube  area  and  number  of  tubes,  440. 


1XDEX. 


655 


Table  6~>.  Distance  of  rivets  from  edge  of  boiler  plate, 
M0, 

"     (Hi.   Diameter  of  rivets  in  boilers,  450. 

"     67.   1'ilcli  of  rivets  for  single-riveted  lap  joints,  453. 

"    68.  Pitch  of  rivets  for  single-riveted  lap  joints,  454. 

"  69.  Pitch  of  rivets  for  double-riveted  lap  joints, 
455. 

"  70.  Pitch  of  rivets  for  double-riveted  lap  joints, 
4.-><i. 

"    71.  Efficiency  of  single-riveted  lap  joints,.  458. 

"    72.  Efficiency  of  single-riveted  lap  joints,  459. 

"    73.  Calculated  diameters  of  stacks,  503. 

"     74.  Diameters  of  stacks  in  actual  practice,  503. 

"  75.  Diameters  of  orifices  in  double  exhaust  nozzles, 
508. 

"  76.  Diameters  of  orifices  in  single  exhaust  nozzles, 
508. 

"  77.  Sizes  of  orifices  in  exhaust  nozzles  in  actual 
service,  509. 

"  78.  Length  of  ordinates  for  the  construction  of 
bells,  515. 

"  79.  Diameter  of  circle  on  ordinates  for  the  con- 
struction of  bells,  515. 

"  80.  Diameter  of  circle  on  ordinates  for  the  con- 
struction of  bells,  516. 

"  81.  Dimensions  of  engine-truck  journals  in  actual 
service,  549. 

"  82.  Dimensions  of  engine-track  journals  computed, 
550. 

"  83.  Dimensions  of  engine-truck  journals  for  con- 
solidation engine**,  551. 

"    84.  Hyperbolic  logarithms,  601. 

Tanks,  capacity  of,  567. 

Tanks  for  tenders,  567. 

Tank  sheets,  thickness  of,  567. 

Tank  valve,  valve  seat,  and  attachments,  567. 

Taper  at  ends  of  piston-rod,  145. 

Tapered  guides,  167. 

Tapered  smoke-stacks,  499. 

Taper  of  bolts  in  main-  and  side-rods,  267. 

Taper  of  bolts  in  side-rods,  299. 

Taper  of  crosshead-pins,  176. 

Taper  of  engine-frame  bolts,  201. 

Taper  of  holes  in  rocker-arms,  74. 

Taper  of  keys  for  main-  and  side-rods,  270,  300. 

Taper  of  legs  in  engine  pedestals,  190. 

Taper  of  link-block  pins,  100. 

Taper  of  mud  plugs  in  boilers,  417. 

Taper  of  piston-rod  keys,  145. 

Taper  of  tread  of  driving  wheels,  263. 

Taper  on  end  of  throttle-valve  rod,  353. 

Temperature  falls  with  the  expansion  of  steam,  617. 

Temperature  of  air  admitted  through  hollow  brick  arch, 
477. 

Templet  for  segments,  form  of,  246. 

Tenacity  of  plates  reduced  by  punching,  451. 

Tender-  and  engine-truck  axles,  difference  between  press- 
ure on,  582. 

Tender  draw-heads,  565. 

Tender-frame,  iron,  569. 

Tender-frames,  longitudinal  bolts  through,  563. 

Tender-frames,  wooden,  565. 

Tenders.  563. 

Tenders,  points  of  support  under  tank,  563. 

Tenders,  weight  of,  581. 

Tenders  with  water-scoops,  572. 

Tender-truck  axle-box  brasses,  575. 

Tender-truck  axle-box  brasses,  width  of,  581. 

Tender-truck  axle-box  brasses  with  lead  lining,  577. 

Tender-truck  axle-boxes,  57."). 

Tender-truck  axle-boxes,  cotton  or  woolen  waste  in,  578. 

Tender-truck  axle-boxes,  covers,  578. 

Tender-truck  axle-boxes,  design  of,  578. 

Tender-truck  axle-boxes,  dust  guard  for,  57*. 

Tender-trnc-k  axle-boxes  for  axles  without  eollars,  f>79. 

Tender-truck  axle-boxes  for  pedestals,  579. 
Tender-truck  axle-boxes,  standard,  579. 


Tender-truck  axle-boxes,  types  of,  579. 
Tender-truck  axle-boxes,  without  wedges,  578. 
Tender-truck  axle-box  wedges,  575. 
Tender-truck  axle-journals,  580. 
Tender-truck  axle-journals,  diameter  of,  581. 
Tender-truck  axle-journals,  pressure  on,  581. 
Tender-truck  axle-journals,  to  compute  dimensions  of, 

582. 
Tender-truck  axle-journals,  to  compute  load  they  can 

bear,  582. 

Tender-truck  axles,  standard,  581. 
Tender-truck  brake  gear,  564,  571. 
Tender-trucks,  571. 
Tender-trucks,  center-bearing,  564. 
Tender-trucks,  safety  chains,  564. 
Tender-trucks,  side-bearing,  564. 
Tensile  force  on  main-rods,  301 . 
Tensile  strength  of  iron  plates,  451,  462. 
Tensile  strength  of  steel  plates,  462. 
Tension  on  driving-wheel  spring  fulcrum,  403. 
Tension  on  driving-wheel  spring  hangers,  403. 
Tension  on  safety-valve  springs,  374,  375,  380. 
Ten-wheeled  engine,  counterbalance  for,  250. 
Ten-wheeled  engine,  driving-axle  boxes  for,  212. 
Ten-wheeled  engine,  driving-wheel  spring  gear  for,  400. 
Ten-wheeled  engine,  pedestals  for,  212. 
Ten-wheeled  engines,  crank-pins  for,  328. 
Ten-wheeled  engines,  distance  between  centers  of  axles, 

188. 

Ten-wheeled  engine,  size  of  driving-axle  journal,  210. 
Ten-wheeled  engines,  knuckle-joint  for  side-rods,  338 
Ten-wheeled  engines,  side-rod  pins  for,  332. 
Ten-wheeled  engines,  side-rods  for,  309. 

Ten-wheeled    locomotive,    computation    of    weight    on 
trucks,  15. 

Ten-wheeled  locomotives,  diameter  of  driving  wheels  of, 
10. 

Ten-wheeled  locomotives,  table  of  weights  and  hauling 
capacity  of,  18. 

Terminal  pressure,  598. 

Terminal  pressure,  computation  of,  599,  601. 

Testing  the  amount  of  piston  clearance,  128. 

Thickness  and  width  of  tires,  265. 

Thickness  of  blades  in  driving-wheel  springs,  415. 

Thickness  of  boiler  lagging,  366. 

Thickness  of  boiler  plates;  formulas  for  computing,  462. 

Thickness  of  boiler  shell,  460. 

Thickness  of  brick  arches,  477. 

Thickness  of  bridges  in  cylinder,  25. 

Thickness  of  bushing  for  eccentric-rod  pins,  103. 

Thickness  of  central  side-rod,  consolidation  engine,  312. 

Thickness  of  counterbalance,  235. 

Thickness  of  crosshead  gibs,  165. 

Thickness  of  cylinder  flanges,  24. 

Thickness  of  cylinder  heads,  24. 

Thickness  of  cylinder  walls,  24. 

Thickness  of  cylinder  walls,  computation  of,  25. 

Thickness  of  cylinder  walls  in  ferry-boats,  25. 

Thickness  of  driving-wheel  equalizing  levers,  402,  408. 

Thickness  of  dry  pipe  for  throttle-valve,  346. 

Thickness  of  engine-frame  pedestal  caps.  194. 

Thickness  of  engine-frame  pedestal  legs,  193. 

Thickness  of  equalizing  lever  in   four-wheeled   engine 
tracks,  538. 

Thickness  of  flanges  on  wedges  for  engine  frames,  186. 

Thickness  of  flue  sheets,  434. 

Thickness  of  furnace  sheets,  434. 

Thickness  of  guides,  166. 

Thickness  of  keys  for  main-  and  side-rods,  300. 

Thickness  of  lead  counterbalance,  251. 

Thickness  of  liners  for  main-  and  side-rods,  270. 

Thickness  of  link,  105. 

Thickness  of  main-rods,  301,  303. 

Thickness  of  main-rod  straps,  283. 

Thickness  of  metal  around  exhaust  passage,  25. 

Thickness  of  metal  around  steam  passage,  25. 

Thickness  of  metal  in  rod  brasses,  317. 

Thickness  of  metal  in  sides  of  saddle,  25. 


656 


INDEX. 


Thickness  of  metal  in  slide-valve,  59. 

Thickness  of  metal  outside  of  slots  through  equalizing 

lever,  406. 
Thickness  of  pipe  connecting  throttle-valve  stuffing-box 

to  dome,  361. 

Thickness  of  plates  in  wrought-iron  ash-pans,  493. 
Thickness  of  reach-rod  pin  bushing,  106. 
Thickness  of  reverse-lever,  109. 
Thickness  of  rocker-arms,  77. 
Thickness  of  segments  of  counterbalance,  243. 
Thickness  of  side-rods  for  eight-wheeled  engines,  309. 
Thickness  of  side-rods  for  Mogul  engines,  310. 
Thickness  of  side-rod  straps,  consolidation  engine,  295. 
Thickness  of  side-rod   straps,  eight-wheeled  passenger 

engines,  286. 

Thickness  of  side-rod  straps,  least,  294. 
Thickness  of  side-rod  straps,  Mogul  engines,  292. 
Thickness  of  slab  frame  braces,  least,  197. 
Thickness  of  slide-valves,  34. 
Thickness  of  stayed  sheets,  481. 
Thickness  of  steam-pipes,  33. 

Thickness  of  two  segments  in  a  counterbalance,  237. 
Thickness  of  wedges  for  engine  frames,  186. 
Thimbles  for  crown  bars,  465. 
Thimbles  for  exhaust  pipes,  505. 
Thimbles  for  frame  pedestals,  182. 

Throttle-lever   handle   and  reverse   lever,    distance  be- 
tween, 356. 

Throttle  lever,  position  of,  355. 
Throttle-lever    quadrants,   to  determine   curvature   and 

length  of,  357. 

Throttle  pipe  and  dry  pipe,  ball-joint  between,  347. 
Throttle  pipe,  construction  of,  344. 
Throttle  pipes,  cross-sectional  area  of  rectangular  part, 

349. 

Throttle  pipes,  diameter  of,  349. 
Throttle  pipes,  thickness  of,  349. 
Throttle  pipe  with  relief-valve,  345. 
Throttle  pipe  with  separator,  351. 
Throttle-valve  bell-crank,  343. 
Throttle-valve  connections,  352. 
Throttle  valve,  diameter  of,  349. 
Throttle-valve  gear  on  top  of  boiler,  359,  364,  366. 
Throttle-valve  gear  with  rod  through  end  of  boiler,  353. 
Throttle-valve  gear  with  rod  through  side  of  dome,  359. 
Throttle-valve  lever,  352,  354,  356. 
Throttle  valve,  lift  of,  358. 
Throttle-valve  quadrant,  353,  354,  355. 
Throttle-valve  rod,  343,  353. 
Throttle-valve  rod,  diameter  of,  355. 
Throttle-valve  rod  gland,  355. 
Throttle-valve  rod  jaw,  353,  362. 
Throttle-valve  rod  jaw,  length  of  circular  arc  on,  363. 
Throttle-valve  rod  stuffing-box,  352,  355. 
Throttle-valve  rod  through  back  head  of  boiler,  352. 
Throttle-valve  rod  with  brass  casing,  355. 
Throttle  valves  and  pipes,  343,  347. 
Throttle-valve  stem,  343. 

Throttle  valve,  to  obtain  a  steam-tight  joint,  345. 
Throw  of  eccentric,  37,  82. 
Throw  of  eccentrics,  amount  of,  104. 
Thrust  against  guides,  computation  of,  159. 
Thrust  against  guides  for  any  position  of  connecting-rod, 

162. 
Thrust  of  connecting-rod,  practical  example  in  finding, 

158. 

Thrust  of  connecting-rod,  to  find,  156. 
Tie  rods  in  boilers,  469. 
Tire  for  driving  wheels,  215. 
Tires  bolted  to  rim  of  wheels,  260. 
Tires,  depth  of  flanges  of,  263. 
Tires,  distance  between  backs  of  flanges  of,  266. 
Tires,  distribution  of  flanged,  261. 
Tires  for  driving  wheels,  259. 

Tires  for  driving  wheels,  shrinkage  allowance,  217. 
Tires,  limit  of  wear  of,  265. 
Tires,  manner  of  fastening  on  wheels,  215. 
Tires,  plain,  261. 


Tires,  shimming  pieces  under,  259. 

Tires  with  a  projection  on  the  inside,  259. 

Tops  for  domes,  380. 

Tops  for  domes  made  in  two  pieces,  387. 

Total  wheel  base,  5. 

Total  wheel  base,  limit  of,  large  engines,  189. 

Total  wheel  base  of  small  engines,  189. 

T-pipe  and  dry-pipe  connections,  349. 

T-pipe  and  dry  pipe,  ball-joint  between,  350. 

Tractive  force,  to  compute,  595. 

Tractive  power  depends  on  diameter  of  cylinders,  stroke, 
diameter  of  drivers,  mean  effective  pressure,  11. 

Tractive  power,  method  of  finding,  11. 

Tractive  power  or  force,  11. 

Tractive  power,  to  compute,  14,  16,  17. 

Train  resistance,  5. 

Transverse  braces  in  boilers,  467. 

Transverse  force  on  main-rods,  301. 

Transverse  key  in  front  end  of  main-rod,  300. 

Transverse  stress  in  boiler  shell,  447. 

Travel  and  lap  of  valve,  to  find,  56. 

Travel  of  valve,  37,  82. 

Travel  of  valve  affects  point  of  cut-off,  54. 

Travel  of  valve,  center  of,  important  position,  44. 

Travel  of  valve  influenced  by  rocker-arms  and  link,  37. 

Travel  of  valve,  to  compute,  40. 

Travel  of  valve  with  lap,  41. 

Tread  of  driving  wheels,  form  of,  263. 

Tread  of  driving  wheels,  standard  form  of,  265. 

Triangle  of  forces  for  finding  thrust  of  connecting-rod, 
158. 

T-ring  for  pistons,  140,  143. 

Trucks,  computation  of  weight  on  engine,  15. 

Truck  wheels,  Allen  Paper  Wheel  Co.,  584. 

Truck  wheels,  Boies  Steel  Wheel  Co.,  585. 

Truck  wheels,  cast-iron,  Pennsylvania  R.  R.,  583. 

Truck  wheels,  difference  between  engine  and  tender,  584. 

Truck  wheels,  S.  M.  Vauclain's  patent,  585. 

Tubes,  arrangement  of,  442. 

Tubes,  cross-sectional  area  of,  438. 

Tubes  for  supporting  brick  arches,  477. 

Tubes,  heating  surface  affected  by  diameter  of,  441. 

Tube  sheet,  manner  of  staying  front,  469. 

Tubes,  length  of,  442. 

Tubes,  number  of,  440. 

Tubes,  ratio  of  diameter  to  length  of,  441. 

Tubes,  space  between,  443. 

Tubes,  thickness,  length,  and  diameter  of  boiler,  437. 

Two-cylinder  compounds,  adjustment  of  valve  gear  in, 
627. 

Two-cylinder  compounds,  diameter  of  cylinders  for,  628. 

Two-cylinder  compound  engines,  620. 

Two-wheeled  engine  truck,  540. 

Two-wheeled  engine  truck,  clearance  between  wheels 
and  cylinder  heads,  542. 

Two-wheeled  engine  truck,  equalizing  lever  for,  546. 

Two-wheeled  engine  truck,  length  of,  542. 

Two-wheeled  engine  truck,  thickness  and  depth  of  equal- 
izing lever,  547. 

Two-wheeled  truck,  computation  of  length  of,  544. 

Two-wheeled  truck,  frame  for,  542. 

Two-wheeled  truck,  graphical  method  of  finding  length 
of,  542. 

Type  of  locomotive,  data  required  for  choice  of,  5. 

Types  of  locomotives,  3,  585. 

Types  of  side-rods,  306. 

TJ 

Units  of  heat  in  coal,  424. 
Upper  engine  frame  brace,  depth  of,  193. 
Useful  rules,  formulas,  and  data,  595. 
Use  of  sand,  511. 


Vacuum  line,  or  zero  line,  597,  607. 
Vacuum  line,  or  zero  line,  location  of,  l 
Vacuum  valves  in  steam  chests,  67. 


657 


Valve,  center  of  its  travel  is  an  important  position,  44. 

Valve,  check,  391. 

Valve  connections,  throttle,  o">_. 

Valve,  diameter  of  throttle,  349. 

Valve  gear,  adjustment  ill  two-cylinder  compounds,  627. 

Valve  gear,  renter  line  of  motion  of,  43. 

Valve  gear,  complete  locomotive,  35. 

Valve  gear,  laying  out.  118. 

Valve  gear,  position  of  axle  for  laying  out,  119. 

Valve  gear,  practical  example  of  setting,  125. 

Valve  gear,  practical  method  of  setting  a  simple,  96. 

Valve  gear,  principles  to  lie  remembered  in  laying  out  a, 

99. 

Valve  gear,  simple,  36. 
Valve  gears  with  rockers,  111. 

Valve  gear,  to  find  the  center  line  of  motion  of,  112. 
Valve,  its  motion  will  depend  on  position  of  saddle-pin, 

lifting-shaft,  and  lengths  of  anus,  119. 
Valve,  load,  42. 
Valve,  linear  advance  of,  45. 
Valve  motion,  influence  of  rocker  on,  36. 
Valve   motions,    infinite   lengths  of  eccentric-rods  give 

symmetrical,  .'!". 

Valve  motion  with  eccentrics  not  on  the  main  axle,  136. 
Valve,  piston,  for  four-cylinder  compounds,  629. 
Valve,  purpose  of  giving  lap  to  glide,  49. 
Valve,  relief,  on  throttle  pipes,  345. 
Valve-rod  end,  bushing  for,  79. 
Valve-rod  end  with  brasses  and  keys,  79. 
Valve-rod  gland,  form  of,  179. 
Valve-rod,  glands  for,  178. 
Valve-rod  joint,  76. 
Valve-rod,  knuckle  joint  in,  79. 
Valve-rod,  springing  of,  76. 
Valves,  balanced  slide,  66. 
Valves,  balanced  slide,  springs  for,  66. 
Valve  seat,  21. 
Valve  seat,  height  of,  29. 
Valve  seat,  length  of,  30. 
Valves  for  pumps,  lift  of,  389. 
Valves,  inside  clearance  of,  in  compounds,  618. 
Valves,  intercepting,  in  compounds,  A.  J.  Pitkin,  620. 
Valves,  intercepting,   in  compounds,  C.  H.  Batchellor, 

625. 

Valve,  slide,  Allen,  68. 
Valve,  slide,  clearance  of,  59. 
Valve,  slide,  conditions  it  must  fulfill,  38. 
Valve,  slide,  friction  of,  64. 
Valve,  slide,  inside  lap  of,  59. 
Valve,  slide,  inside  lead,  59. 
Valve,  slide,  power  required  to  work  a,  63. 
Valve,  slide,  practical  construction  of,  57. 
Valve,  slide,  pressure  on,  64,  65. 
Valve,  slide,  primitive  form  of,  48. 
Valve,  slide,  problems  relating  to  lap  of,  52. 
Valve,  slide,  thickness  of  metal,  59. 
Valve,  slide,  to  find  lap  and  travel  of,  56. 
Valve,  slide,  to  find  lap  of,  55. 
Valve,  slide,  travel  of,  affects  point  of  cut-off,  54. 
Valves,  roller  isliile.  il.~>. 
Valves,  slide,  34. 

Valves,  slide,  classification  of,  39. 
Valves,  slide,  duty  of,  3f>. 
Valves,  slide,  thickness  of  metal,  34. 
Valve,  throttle.  :;t:i. 
Valve  travel.  37.  M>. 

Valve  travel  influenced  liy  rocker-arms  and  link.  l!7. 
Valve,  travel  of,  with  lap,  40. 
Valve,  travel,  to  compute,  40. 
Valve  with  lead,  position  of  eccentric,  43. 
Variable  motion  of  piston,  cause  of,  50. 
Variation  of  pressure  in  cylinders  in  compound  engines, 

017. 

Vaiidain,  S.  M.,  four-cylinder  compounds,  629. 
Vauclain,    S.    M..  starting   valve    for   four-cylinder  com- 
pounds. 631. 

Vertical  clearance  in  engine  pedestals  for  driving  boxes, 
189. 


Vertical  movement  of  driving  boxes,  189. 
Volume  of  receiver  in  compounds,  629. 
Volute  springs  for  lifting-shaft  counterbalance,  107. 
Volute  springs  for  lifting-shaft  counterbalance,  dimen- 
sions of,  108. 

W 

Walls,  thickness  of  cylinder,  24. 

Washers  for  crown  bars,  465. 

Water-grate  bearers,  489. 

Water  grates,  486. 

Water  grates,  inclination  of,  486. 

Water  grates,  manner  of  fastening  in  furnace,  489. 

Water-gauge  glass,  562. 

Water-scoops  for  tenders,  572. 

Water  space,  width  of,  434. 

Weakest  part  of  piston-rods,  148. 

Wear  and  re-planing  of  guides,  allowance  for,  167,  168. 

Wear  of  cast-iron  crossheads  and  case-hardened  guides, 
151. 

Wear  of  eccentric-rod  pins,  103. 

Wear  of  tires,  limit  of,  265. 

Wear  of  upper  crosshead  gib,  162. 

Wedges  and  bolts  for  engine-frame  pedestals,  propor- 
tions of,  186. 

Wedges  for  engine-frame  pedestals,  184. 

Wedges  for  tender-truck  axle-boxes,  575. 

Weight  of  counterbalance  to  balance  crank  and  an  addi- 
tional load  on  crank-pin,  233. 

Weight  of  crank-pin  hub,  258. 

Weight  of  crank  referred  to  crank-pin,  228. 

Weight  of  foot-board,  528. 

Weight  of  lead  in  hollow  rims  of  wheel  for  counter- 
balance, 255. 

Weight  of  locomotives,  data  used  for  computing,  8,  15. 

Weight  of  locomotives,  table  of,  17,  18,  19. 

Weight  of  rails  must  be  known  to  determine  the  number 
of  driving  wheels,  8. 

Weight  of  tenders,  581. 

Weight  on  drivers,  to  compute,  8,  14,  595. 

Weight  on  driving  wheels  limited  by  bridges,  8. 

Weight  on  engine  trucks,  computation  of,  15. 

Weights  to  be  counterbalanced,  243. 

Weights,  to  find  common  center  of  gravity  of  any  two,  240. 

Welt  pieces  on  riveted  joints,  459. 

Westinghouse  brake,  manner  of  attaching,  585. 

Wheel  base  of  small  engines,  total,  189. 

Wheel  base,  rigid,  5. 

Wheel  base,  total,  5. 

Wheel  base,  total,  limit  of,  in  large  engines,  189. 

Wheel,  center  of  gravity  of  lead  in  rim  of,  257. 

Wheel  centers,  standard  diameter  of,  217. 

Wheel  covers  for  driving  wheels,  530. 

Wheel  covers  for  engine  trucks,  535. 

Wheel,  cross-sectional  area  of  rim  of.  ""><>. 

Wheel,  effect  of  lead  counterbalance  in  rim  of,  257. 

Wheel  fit  of  crank-pins  for  ten-wheeled,  Mogul,  and  con- 
solidation engines,  3.'!7. 

Wheel  fit  of  crank-pins  in  eight-wheeled  engines,  325. 

Wheel  lits  on  driving-axles,  L'14. 

Wheels,  clearance  between  rails  and  flanges  of,  L'lil. 

Wheels,  difference  between  engine  and  tender  truck,  584. 

Wheels,  dished  driving,  220. 

Wheels,  driving.  2 1.'.. 

Wheels,  driving,  pressure  for  forcing  on  axle.  L'L'O. 

Wheels  for  trucks,  Allen  Paper  Wheel  Co.,  584. 

Wheels  for  trucks,  Boies  Steel  Wheel  Co.,  r,sr>. 

Wheels  for  trucks,  cast-iron,  Pennsylvania  l{.  }{.,  583. 

Wheels  for  trucks.  S.  M.  Vauclain's  patent,  .".--."p. 

Wheels,  tires  for  driving,  L'V.i. 

Wheels  under  hard-coal  burners,  position  of,  1 89. 

Wl Is  with  tires  bolted  to  them,  260. 

Whistle,  381. 

Whistle,  chime,  383. 

Whistle  levers,  383. 

Whistle,  position  of,  380. 

Width  and  thickness  of  flanged  tires.  265. 

Width  of  brass  packing  rings  for  pistons,  144. 


658 


INDEX. 


Width  of  crossheads,  165. 

Width  of  driving-axle  box,  208. 

Widtli  of  engine  frames,  191. 

Width  of  engine  pedestal  openings,  190. 

Width  of  exhaust  ports,  29. 

Width  of  fire-box,  greatest,  427. 

Width  of  flanges  on  driving-axle  boxes,  204. 

Width  of  keys  for  main-  and  side-rods,  300. 

Width  of  piston  packing,  139. 

Width  of  plain  tires,  261. 

Width  of  rocker-arm,  77. 

Width  of  steam  ports,  29,  30. 

Width  of  tender-truck  axle-box  brasses,  581. 

Wings  on  crosshead,  149. 

Wings  with  glass  disks  on  crossheads,  151. 

Wood-burning  engines,  position  of  fire-box  in,  188. 

Wood-burning  fire-boxes,  430. 

Wood-burning  stack,  497. 


Wooden  floor  on  top  of  fool-plates,  528. 

Wooden  pilots,  517. 

Wooden  tender  frames,  565. 

Woolen  or  cotton  waste  in  tender-truck  axle-boxes,  578. 

Wootten  boiler,  478. 

Work  and  energy,  612. 

Work  done  by  a  locomotive,  12. 

Wrought-iron  crank-pins,  317. 

Wrought-irou  draw-head  on  bumper  beam,  517. 


Yoke  for  dry  pipe,  347. 
Yokes  for  bells,  513. 


Z 


Zero  line,  or  line  of  perfect  vacuum,  597,  607. 

Zero  line,  or  line  of  perfect  vacuum,  location  of,  602. 


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